Linear Cross-Head Bearing for Stirling Engine

ABSTRACT

An external combustion engine including a burner element, a heater head, a piston cylinder containing a piston, a cooler and a crankcase. The crankcase includes a crankshaft, a piston rod connected to the piston, a drive mechanism for converting the linear motion of the piston rod to rotary motion of the crankshaft and a linear cross-head bearing that is connected rigidly to the piston rod at one end and to the drive mechanism at the other end. Also the external combustion engine includes a piston clearance seal and a piston rod seal unit that has floating rod seals. The piston includes a inner dome to reduce axial heat transfer via radiation and convection.

CROSS REFERENCE TO RELATED APPLICATIONS

The present application is a Continuation-In-Part of U.S. patentapplication Ser. No. 14/553,824, filed Nov. 11, 2015, and entitledAnnular Venturi Burner for Stirling Engine (Attorney Docket No. 195),which is a Continuation-In-Part of U.S. Patent Application Ser. No.61/908,468, filed Nov. 25, 2013, and entitled Annular Venturi Burner forStirling Engine (Attorney Docket No. 190), which is also aContinuation-In-Part of U.S. patent application Ser. No. 14/211,621,filed Mar. 14, 2014, and entitled Stirling Cycle Machine (AttorneyDocket No. 192), which is a Continuation-In-Part of U.S. patentapplication Ser. No. 13/836,946, filed Mar. 15, 2013, and entitledStirling Cycle Machine (Attorney Docket No. 187), which is aContinuation-In-Part of U.S. patent application Ser. No. 13/447,990,filed Apr. 16, 2012, and entitled Stirling Cycle Machine (AttorneyDocket No. 184), which claims priority to:

U.S. Provisional Patent Application Ser. No. 61/476,180, filed Apr. 15,2011 and entitled Stirling Cycle Machine (Attorney Docket No. 181); andU.S. Provisional Patent Application Ser. No. 61/482,897, filed May 5,2011 and entitled Stirling Cycle Machine (Attorney Docket No. 182), allof which are hereby incorporated herein by reference in theirentireties.

U.S. patent application Ser. No. 13/447,990, filed Apr. 16, 2012, andentitled Stirling Cycle Machine (Attorney Docket No. 184), is also aContinuation-In-Part application of U.S. patent application Ser. No.12/829,320, filed Jul. 1, 2010 and entitled Stirling Cycle Machine, nowU.S. Publication No. US-2011-0011078-A1, published Jan. 20, 2011(Attorney Docket No. 178), which claims priority to U.S. ProvisionalPatent Application Ser. No. 61/222,361, filed Jul. 1, 2009 and entitledStirling Cycle Machine (Attorney Docket No. 175), both of which arehereby incorporated herein by reference in their entireties.

TECHNICAL FIELD

The present invention relates to machines and more particularly, to aStirling cycle machine and components thereof.

BACKGROUND INFORMATION

Many machines, such as internal combustion engines, external combustionengines, compressors, and other reciprocating machines, employ anarrangement of pistons and drive mechanisms to convert the linear motionof a reciprocating piston to rotary motion. In most applications, thepistons are housed in a cylinder. A common problem encountered with suchmachines is that of friction generated by a sliding piston resultingfrom misalignment of the piston in the cylinder and lateral forcesexerted on the piston by linkage of the piston to a rotating crankshaft.These increased side loads increase engine noise, increase piston wear,and decrease the efficiency and life of the engine. Additionally,because of the side loads, the drive requires more power to overcomethese frictional forces, thus reducing the efficiency of the machine.

Improvements have been made on drive mechanisms in an attempt to reducethese side loads, however, many of the improvements have resulted inheavier and bulkier machines.

Accordingly, there is a need for practical machines with minimal sideloads on pistons.

SUMMARY

In accordance with one aspect of the present invention, an externalcombustion engine is disclosed. The external combustion enginecontaining a working fluid and includes a burner element for heating theworking fluid of the engine, at least one heater head defining a workingspace containing the working fluid, at least one piston cylindercontaining a piston for compressing the working fluid, a cooler forcooling the working fluid, a crankcase including a crankshaft forproducing an engine output, a rocking beam rotating about a rocker pivotfor driving the crankshaft, a piston rod connected to the piston, arocking beam driven by the piston rod, and a connecting rod connected ata first end to the rocking beam and at a second end to a crankshaft toconvert rotary motion of the rocking beam to rotary motion of thecrankshaft. Also, the external combustion engine including a piston rodseal unit including a housing, a cylinder gland, and at least onefloating rod seal assembly mounted in the cylinder gland, the floatingrod seal assembly comprising at least one rod seal mounted onto thefloating rod seal assembly.

Some embodiments of this aspect of the present invention include one ormore of the following. Where the piston rod seal unit further includes ascraper ring. Where the piston rod seal unit further includes a particletrap. Wherein the piston rod seal unit further includes a port. Whereinthe piston rod seal unit further includes a filter. Wherein the floatingrod seal assembly further includes an outer ring, and at least onebushing. Wherein the piston rod seal unit further includes wherein therod seal is a spring energized seal.

In accordance with one aspect of the present invention, a piston rodseal unit is disclosed. The piston rod seal unit includes a housing, acylinder gland, and at least one floating rod seal assembly mounted inthe cylinder gland, the floating rod seal assembly comprising at leastone rod seal mounted onto the floating rod seal assembly.

Some embodiments of this aspect of the present invention include one ormore of the following. Wherein the piston rod seal unit further includesa scraper ring. Wherein the piston rod seal unit further includes aparticle trap. Wherein the piston rod seal unit further includes a port.Wherein the piston rod seal unit further includes a filter. Wherein thefloating rod seal assembly further includes an outer ring, and at leastone bushing. Wherein the piston rod seal unit further includes whereinthe rod seal is a spring energized seal.

In accordance with one aspect of the present invention, an externalcombustion engine is disclosed. The external combustion enginecontaining a working fluid and including a piston rod seal unitincluding a housing, a cylinder gland, and at least one floating rodseal assembly mounted in the cylinder gland, the floating rod sealassembly including at least one rod seal mounted onto the floating rodseal assembly, and an airlock space separating a crankcase and a workingspace for maintaining a pressure differential between a crankcasehousing and a working space housing.

Some embodiments of this aspect of the present invention include one ormore of the following. Wherein the airlock pressure regulator is abidirectional pressure regulator for maintaining a predeterminedpressure differential between the crankcase and one of the airlock spaceand working space. Wherein the airlock pressure regulator includes afilter, a compressor, a pressure regulating spool valve, and a linearposition sensor, wherein the linear position sensor produces a signalindicative of the regulating spool valve position. Wherein the airlockpressure regulator further includes a controller. Wherein the linearposition sensor is an LVDT. Wherein the controller uses the linearposition sensor to regulate a pump speed.

In accordance with one aspect of the present invention, a rod sealassembly is disclosed. The rod seal assembly includes a housing betweentwo spaces configured to receive a reciprocating rod, the reciprocatingrod disposed within a first space and a second space, a floating bushingconfigured to move axially and radially within the housing and disposedcoaxially around the reciprocating rod, a rod seal configured to sealthe outside diameter of the reciprocating rod relative to an insidesurface of the floating bushing, and at least one stationary bushingfixed within the housing that may form a seal with the floating bushingto the axial flow of fluid in the presence of a pressure differencebetween the two spaces.

Some embodiments of this aspect of the present invention include one ormore of the following. Wherein the floating bushing is configured tomove radially to center on the piston rod when the pressure differencebetween the first and second space is small and form the seal with thestationary bushing when the pressure difference is larger. Wherein therod seal is a spring energized seal. Wherein the floating bushingfurther comprises a circumferential flange on the outside surface thatis configured to extend into the annular space and form a seal with oneof the stationary bushings. Wherein the rod seal is formed of a PTFEcomposite. Wherein the floating bushing and stationary bushing areformed of a wear resistance metal. Wherein the assembly further includesa scraper ring disposed coaxially around the piston rod and disposedwithin housing between the floating seal and the first space, and apassage connecting the first space to an annular gap disposed around thereciprocating rod between the scraper ring and the floating seal.Wherein the assembly further includes a magnetic particle trap disposedbetween the scraper ring and floating seal.

In accordance with one aspect of the present invention, a rod sealassembly is disclosed. The rod seal assembly includes a housing betweentwo spaces configured to receive a reciprocating rod, the reciprocatingrod disposed between a first space and a second space, a floatingclearance bushing configured to move axially and radially within thehousing and disposed coaxially around the reciprocating rod and forms aclearance seal with the reciprocating rod, and at least one stationaryannular element fixed within the housing configured to form a face sealwith the floating clearance bushing.

Some embodiments of this aspect of the present invention include one ormore of the following. Wherein the floating clearance bushing isconfigured to move radially to center on the piston rod when thepressure difference between the first and second space is small and formthe seal with the stationary annular element when the pressuredifference is larger. Wherein the assembly further includes a springenergized face seal on at least one end of the floating clearancebushing. Wherein the assembly further includes a second floatingclearance bushing disposed around the reciprocating rod, and two springenergized lip seals disposed around the reciprocating piston rod andaxially located within or between the two floating clearance bushings.Wherein the assembly further includes a scraper ring disposed coaxiallyaround the piston rod and disposed within housing between the floatingseal and the first space, and a passage connecting the first space to anannular gap disposed around the reciprocating rod between the scraperring and the floating seal. Wherein the assembly further includes amagnetic particle trap disposed between the scraper ring and floatingseal.

In accordance with one aspect of the present invention, a floating rodseal is disclosed. The floating seal includes a rod seal attached to afloating bushing, wherein the rod seal forms a leak tight joint with afloating bushing, two annular shaped stationary bushings that arelocated approximately coaxially with respect to the rod seal and placedon the inside diameter of an outer ring such that the two ends of thestationary bushings form an axial gap, and the floating bushing whichincludes an inner surface to seal to the rod seal and a circumferentialrib on the outside surface, where the circumferential rib is captured inthe axial gap and may move radially within the outer ring and may form aseal with one bushing.

In accordance with one aspect of the present invention, a floating rodseal is disclosed. The floating seal includes a floating clearancebushing that forms a clearance seal with the piston rod and floatswithin cylinder-gland housing located between a workspace the air lock.The floating clearance bushing moves radially when the pressuredifference between the workspace and the air lock are minimal and formsa seal with a fixed annular section of the housing when the pressuredifference is large.

In accordance with one aspect of the present invention, an externalcombustion engine is disclosed containing a working fluid and comprisinga burner element for heating the working fluid of the engine, at leastone heater head defining a working space containing the working fluid,at least one piston cylinder containing a piston for compressing theworking fluid, at least one cooler for cooling the working fluid and acrankcase. The crankcase comprises a crankshaft for producing an engineoutput, a piston rod connected to the piston, a drive mechanism thatconverts the linear motion of the piston rod to rotary motion of thecrankshaft, and a linear cross-head bearing comprising a journal and aguide. One end of the journal bearing is rigidly attached to the pistonrod and the other end is attached to the drive mechanism. The guide islocated outside the working space and the linear cross-head bearingsolely constrains the motion of the piston.

Some embodiments of this aspect of the present invention include one ormore of the following: the ratio of the linear cross-head bearing lengthover diameter is greater than 2.0; the linear cross-head bearing is ahydrodynamic bearing supplied with pressurized oil from an annulus onthe guide; a piston rod seal unit including a housing; a floatingclearance bushing configured to move axially and radially within thehousing and disposed coaxially around the piston rod and at least onestationary annular element fixed within the housing configured to form aface seal with the floating clearance bushing; a piston rod seal unitincluding a housing, a floating rod seal assembly with at least one rodseal mounted onto the floating rod seal assembly. a piston rod seal unitincluding a housing, a cylinder gland, and at least one floating rodseal assembly mounted in the cylinder gland where a rod seal is mountedonto the floating rod seal assembly; the rod seal is an spring energizedseal; the piston includes a clearance seal where the clearance seal mayinclude radial grooves on the outside diameter and O-rings on the insidediameter and the clearance seal may include at least one of thefollowing materials; graphite, PTFE, UHMWPE, antimony.

In accordance with one aspect of the present invention, an externalcombustion engine is disclosed containing a working fluid and comprisinga burner element for heating the working fluid of the engine, at leastone heater head defining a working space containing the working fluid,at least one piston cylinder containing a piston for compressing theworking fluid, at least one cooler for cooling the working fluid and acrankcase. The crankcase comprises a crankshaft for producing an engineoutput, a piston rod connected to the piston, a rocking beam rotatingabout a rocker pivot for driving the crankshaft, a rocking beam drivenby the piston rod, and a connecting rod connected at a first end to therocking beam and at a second end to a crankshaft to convert rotarymotion of the rocking beam to rotary motion of the crankshaft and alinear cross-head bearing comprising a journal and a guide. One end ofthe journal bearing is rigidly attached to the piston rod and the otherend is attached to the drive mechanism. The guide is located outside theworking space and the linear cross-head bearing solely constrains themotion of the piston.

Some embodiments of this aspect of the present invention include one ormore of the following: the ratio of the linear cross-head bearing lengthover diameter is greater than 2.0; the linear cross-head bearing is ahydrodynamic bearing supplied with pressurized oil from an annulus onthe guide; a piston rod seal unit including a housing; a floatingclearance bushing configured to move axially and radially within thehousing and disposed coaxially around the piston rod and at least onestationary annular element fixed within the housing configured to form aface seal with the floating clearance bushing; a piston rod seal unitincluding a housing, a floating rod seal assembly with at least one rodseal mounted onto the floating rod seal assembly. a piston rod seal unitincluding a housing, a cylinder gland, and at least one floating rodseal assembly mounted in the cylinder gland where a rod seal is mountedonto the floating rod seal assembly; the rod seal is an spring energizedseal; the piston includes a clearance seal where the clearance seal mayinclude radial grooves on the outside diameter and O-rings on the insidediameter and the clearance seal may include at least one of thefollowing materials; graphite, PTFE, UHMWPE, antimony.

In accordance with one aspect of the present invention, an externalcombustion engine is disclosed containing a working fluid and comprisinga burner element for heating the working fluid of the engine, at leastone heater head defining a working space containing the working fluid,at least one piston cylinder containing a piston for compressing theworking fluid, at least one cooler for cooling the working fluid and acrankcase. The crankcase comprises a crankshaft for producing an engineoutput, a piston rod connected to the piston, a drive mechanism thatconverts the linear motion of the piston rod to rotary motion of thecrankshaft, and a piston comprising a piston base with a seal, an outershell that mounts on the base and an inner shell that is shorter andnarrower than the outer shell and defines with the piston base a volumeabove the inner surface of the piston base.

Some embodiments of this aspect of the present invention include one ormore of the following: a small orifice in the piston base, an orifice inthe piston base with a diameter between 0.002 and 0.008 inches, and portin the inner shell. In accordance with one aspect of the presentinvention, an external combustion engine is disclosed. The externalcombustion engine containing a working fluid and including a burnerelement for heating the working fluid of the engine, at least one heaterhead defining a working space containing the working fluid, at least onepiston cylinder containing a piston for compressing the working fluid, acooler for cooling the working fluid, a crankcase. The crankcaseincludes a crankshaft for producing an engine output, a rocking beamrotating about a rocker pivot for driving the crankshaft, a piston rodconnected to the piston, a rocking beam driven by the piston rod, and aconnecting rod connected at a first end to the rocking beam and at asecond end to a crankshaft to convert rotary motion of the rocking beamto rotary motion of the crankshaft. The external combustion engine alsoincludes an airlock space separating the crankcase and the working spacefor maintaining a pressure differential between the crankcase housingand the working space housing and an airlock pressure regulatorconnected between the crankcase and one of the airlock space and workingspace.

Some embodiments of this aspect of the present invention include one ormore of the following. A first seal for sealing the crankcase from theairlock space, wherein the seal is a rolling diaphragm. A seal forsealing the workspace from the airlock space, wherein the seal is a pairof oppositely disposed rolling diaphragms. A second seal for sealing theworkspace from the airlock space, wherein the seal is a high pressureseal. Wherein the airlock pressure regulator is a bidirectional pressureregulator for maintaining a predetermined pressure differential betweenthe crankcase and one of the airlock space and working space. Whereinthe airlock pressure regulator includes a filter, a compressor, apressure regulating spool valve, and a linear position sensor, whereinthe linear position sensor produces a signal indicative of theregulating spool valve position. Wherein the airlock pressure regulatorfurther includes a controller. Wherein the linear position sensor is anLVDT. Wherein the controller uses the linear position sensor to regulatea pump speed. Wherein the airlock pressure regulator further includes adrain port and a fill port, wherein the drain port and the fill port areselectively connected. Wherein the controller sends a stop command tothe drive based at least in part on the position of the linear positionsensor. Wherein the external combustion engine further includes a burnercontroller for a multi-burner including a master controller and anindividual combustion control circuit. Wherein the master controllercontrols a variable resistance element connected to the fuel line ineach burner in the multi-burner. Wherein the external combustion enginefurther includes an over-temperature circuit, wherein theover-temperature circuit monitors a temperature on each of the heaterhead temperatures and may disable the fuel valve supplying a burnerheating a heater head. Wherein the external combustion engine furtherincludes a flow-detection circuit wherein the flow-detection circuitdisables all the fuel valves when the flow-detection circuit detects aflow that is below a predetermined threshold.

In accordance with one aspect of the present invention, a rocking beamdrive mechanism for a machine is disclosed. The drive mechanism includesa rocking beam having a rocker pivot, at least one cylinder and at leastone piston. The piston is housed within a respective cylinder. Thepiston is capable of substantially linearly reciprocating within therespective cylinder. Also, the drive mechanism includes at least onecoupling assembly having a proximal end and a distal end. The proximalend is connected to the piston and the distal end is connected to therocking beam by an end pivot. The linear motion of the piston isconverted to rotary motion of the rocking beam.

Some embodiments of this aspect of the present invention include one ormore of the following: where the rocking beam is coupled to a crankshaftby way of a connecting rod. In this embodiment, the rotary motion of therocking beam is transferred to the crankshaft. Also, where the cylindermay further include a closed end and an open end. The open end furtherincludes a linear bearing connected to the cylinder. The linear bearingincludes an opening to accommodate the coupling assembly. Also, wherethe coupling assembly further includes a piston rod and a link rod. Thepiston rod and link rod are coupled together by a coupling means. Thecoupling means is located beneath the linear bearing. Also, where thedrive mechanism also includes a seal, where the seal is sealablyconnected to the piston rod. Also, where the seal is a rollingdiaphragm. Also, in some embodiments, the coupling means is a flexiblejoint. In some embodiments, the coupling means is a roller bearing. Insome embodiments, the coupling means is a hinge. In some embodiments,the coupling means is a flexure. In some embodiments, the coupling meansis a journal bearing joint.

In accordance with another aspect of the present invention, a Stirlingcycle machine is disclosed. The machine includes at least one rockingdrive mechanism where the rocking drive mechanism includes: a rockingbeam having a rocker pivot, at least one cylinder and at least onepiston. The piston is housed within a respective cylinder. The piston iscapable of substantially linearly reciprocating within the respectivecylinder. Also, the drive mechanism includes at least one couplingassembly having a proximal end and a distal end. The proximal end isconnected to the piston and the distal end is connected to the rockingbeam by an end pivot. The linear motion of the piston is converted torotary motion of the rocking beam. Also, a crankcase housing the rockingbeam and housing a first portion of the coupling assembly is included. Acrankshaft coupled to the rocking beam by way of a connecting rod isalso included. The rotary motion of the rocking beam is transferred tothe crankshaft. The machine also includes a working space housing the atleast one cylinder, the at least one piston and a second portion of thecoupling assembly. A seal is included for sealing the workspace from thecrankcase.

Some embodiments of this aspect of the present invention include one ormore of the following: where the seal is a rolling diaphragm. Also, thecylinder may further include a closed end and an open end. The open endfurther includes a linear bearing connected to the cylinder. The linearbearing includes an opening to accommodate the coupling assembly. Also,where the coupling assembly further includes a piston rod and a linkrod. The piston rod and link rod are coupled together by a couplingmeans. The coupling means may be located beneath the linear bearing.Also, the machine may also include a lubricating fluid pump in thecrankcase. In some embodiments, the lubricating fluid pump is amechanical lubricating fluid pump driven by a pump drive assembly, thepump drive assembly being connected to and driven by the crankshaft. Insome embodiments, the lubricating fluid pump is an electric lubricatingfluid pump. The machine may also include a motor connected to thecrankshaft. The machine may also include a generator connected to thecrankshaft.

In accordance with another aspect of the present invention, a Stirlingcycle machine is disclosed. The machine includes at least two rockingdrive mechanisms. The rocking drive mechanisms each include a rockingbeam having a rocker pivot, two cylinders, and two pistons. The pistonseach housed within a respective cylinder. The pistons are capable ofsubstantially linearly reciprocating within the respective cylinder.Also, the drive mechanisms include two coupling assemblies having aproximal end and a distal end, the proximal end being connected to thepiston and the distal end being connected to the rocking beam by an endpivot. The linear motion of the piston is converted to rotary motion ofthe rocking beam. The machine also includes a crankcase housing therocking beam and housing a first portion of the coupling assemblies.Also, a crankshaft coupled to the rocking beam by way of a connectingrod. The rotary motion of the rocking beam is transferred to thecrankshaft. The machine also includes a lubricating fluid pump in thecrankcase for pumping lubricating fluid to lubricate the crankshaft andthe rocking beam and the first portion of the coupling assemblies. Also,a working space housing the cylinders, the pistons and the secondportion of the coupling assemblies. A rolling diaphragm for sealing theworkspace from the crankcase is also included.

Some embodiments of this aspect of the present invention include one ormore of the following: where the cylinder may further include a closedend and an open end. The open end further includes a linear bearingconnected to the cylinder. The linear bearing includes an opening toaccommodate the coupling assembly. Also, where the coupling assemblyfurther includes a piston rod and a link rod. The piston rod and linkrod are coupled together by a coupling means. The coupling means may belocated beneath the linear bearing. Also, where the coupling means is aflexible joint. In some embodiments, also disclosed is where thecoupling means is a roller bearing.

Other embodiments of this aspect of the present invention relate to oneor more of an external combustion engine containing a working fluidcomprising a burner element for heating the working fluid of the engine,at least one heater head defining a working space containing the workingfluid, at least one piston cylinder containing a piston for compressingthe working fluid, a cooler for cooling the working fluid, a crankcasecomprising a crankshaft for producing an engine output, a rocking beamrotating about a rocker pivot for driving the crankshaft, a piston rodconnected to the piston, a rocking beam driven by the piston rod, and aconnecting rod connected at a first end to the rocking beam and at asecond end to the crankshaft to convert rotary motion of the rockingbeam to rotary motion of the crankshaft wherein the piston reciprocatesalong a substantially linear piston axis in the crankcase and thecrankshaft is arranged below a limit of the piston axis in thecrankcase.

A still further embodiment of the invention relate to one or moreembodiments of an external combustion engine containing a working fluidcomprising a burner element for heating the working fluid of the engine,at least one heater head defining a working space containing the workingfluid, at least one piston cylinder containing a piston for compressingthe working fluid, a cooler for cooling the working fluid, a crankcasecomprising, a crankshaft for producing an engine output, a rocking beamrotating about a rocker pivot for driving the crankshaft, a piston rodconnected to the piston, a rocking beam driven by the piston rod, aconnecting rod connected at a first end to the rocking beam and at asecond end to a crankshaft to convert rotary motion of the rocking beamto rotary motion of the crankshaft, and an airlock space separating thecrankcase and the working space for maintaining a pressure differentialbetween the crankcase housing and the working space housing.

A still further embodiment of the invention relate to one or moreembodiments of a heating element for heating an external combustionengine or machine comprising a burner element for heating the workingfluid of the engine, a blower providing air or other gas forfacilitating ignition and combustion in the burner, a preheater definingan incoming air passage and an exhaust passage separated by an exhaustmanifold wall for heating incoming air from the hot exhaust expelledfrom the heating element, a fuel injector for supplying fuel to mix withthe incoming air, an igniter to ignite the fuel/air mixture, aprechamber defining an inlet for receiving the fuel/air mixture andpromoting ignition of the mixture, a combustion chamber disposedlinearly below the prechamber for maintaining supporting a flamedeveloped and ignited in the prechamber, an electronic control unit forcontrolling ignition and combustion operations of the burner, andwherein the combustion chamber is connected to the exhaust passage intowhich the exhausted combustion gases are pushed to heat the incoming airfollowing combustion and heating of the engine or machine. These aspectsof the invention are not meant to be exclusive and other features,aspects, and advantages of the present invention will be readilyapparent to those of ordinary skill in the art when read in conjunctionwith the appended claims and accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

These and other features and advantages of the present invention will bebetter understood by reading the following detailed description, takentogether with the drawings wherein:

FIGS. 1A-1E depict the principle of operation of a prior art Stirlingcycle machine;

FIG. 2 shows a view of a rocking beam drive in accordance with oneembodiment;

FIG. 3 shows a view of a rocking beam drive in accordance with oneembodiment;

FIG. 4 shows a view of an engine in accordance with one embodiment;

FIGS. 5A-5D depicts various views of a rocking beam drive in accordancewith one embodiment;

FIG. 6 shows a bearing style rod connector in accordance with oneembodiment;

FIGS. 7A-7B show a flexure in accordance with one embodiment;

FIG. 8 shows the operation of pistons of an engine in accordance withone embodiment;

FIG. 9A shows an unwrapped schematic view of a working space andcylinders in accordance with one embodiment;

FIG. 9B shows a schematic view of a cylinder, heater head, andregenerator in accordance with one embodiment;

FIG. 9C shows a view of a cylinder head in accordance with oneembodiment;

FIG. 10A shows a view of a rolling diaphragm, along with supporting topseal piston and bottom seal piston, in accordance with one embodiment;

FIG. 10B shows an exploded view of a rocking beam driven engine inaccordance with one embodiment;

FIG. 10C shows a view of a cylinder, heater head, regenerator, androlling diaphragm, in accordance with one embodiment;

FIG. 10D shows various views of a rolling diaphragm during operation, inaccordance with one embodiment;

FIG. 11 shows a view of an external combustion engine in accordance withone;

FIGS. 12A-12E show views of various embodiments of a rolling diaphragm;

FIG. 13A shows a schematic of a rolling diaphragm identifying variousload regions;

FIG. 13B shows a schematic of the rolling diaphragm identifying theconvolution region;

FIG. 14 shows a view of a piston and piston seal in accordance with oneembodiment;

FIG. 15 shows a view of a piston rod and piston rod seal in accordancewith one embodiment;

FIGS. 16A-16B show views of a piston guide ring in accordance with oneembodiment;

FIG. 17 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 18 shows a portion of a cross section of a tube heat exchanger inaccordance with one embodiment;

FIG. 19 shows a view of a heater head of an engine in accordance withone embodiment;

FIG. 20A shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 20B shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 21A shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 21B shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 22A shows view of a tube heat exchanger in accordance with oneembodiment;

FIG. 22B shows a view of a tube heat exchanger in accordance with oneembodiment;

FIGS. 23A and 23B show a regenerator of a Stirling cycle engine inaccordance with one embodiment;

FIGS. 24A-24E show various configurations of a regenerator of a Stirlingcycle engine in accordance with various embodiments;

FIGS. 25A-25C show various views of an engine in accordance with severalembodiments;

FIGS. 26A and 26B show views of a cooler for an engine in accordancewith some embodiments;

FIGS. 27A and 27B show a view of a cooler for an engine in accordancewith one embodiment;

FIGS. 27C and 27D show a view of a cooler for an engine in accordancewith one embodiment;

FIGS. 28A-28C show views of an intake manifold for an engine inaccordance with one embodiment;

FIG. 29 is a gaseous fuel burner coupled to a Stirling cycle engine,where the ejector is a venturi, according to one embodiment;

FIG. 30A is the burner of FIG. 29 showing the air and fuel flow paths inaccordance with one embodiment;

FIG. 30B is a graphical representation of the pressure across the burnerin accordance with one embodiment;

FIG. 31 shows a schematic of an embodiment of the burner with automatedfuel control for variable fuel properties;

FIG. 32 shows a schematic of another embodiment of the burner withtemperature sensor and engine speed control loop;

FIG. 33 shows a schematic of yet another embodiment of the burner withtemperature sensor and oxygen sensor control loop;

FIG. 34 shows an alternative embodiment of the ejector wherein the fuelis fed directly into the ejector;

FIG. 35 shows a cross section of an engine in accordance with oneembodiment;

FIGS. 36A and 36B show a cross-sectional view of a Stirling cyclemachine having an inverted rocking beam design in accordance with oneembodiment;

FIGS. 36C-36E show various views of a piston and piston rod assembly inaccordance with one embodiment;

FIG. 37A shows a view of an embodiment of the rocking beam with a conrodbearing ratio of 1.6;

FIG. 37B shows a view of an embodiment of the rocking beam with a conrodbearing ratio of 1.0;

FIG. 38A shows an oil pump according to one embodiment;

FIG. 38B shows a Gerotor displacement pumping unit according to oneembodiment;

FIG. 39 shows an embodiment of a high pressure rod seal;

FIG. 40A shows another embodiment of a high pressure rod seal includinga spring energized lip seal;

FIG. 40B is a hydraulic high pressure piston rod seal set inside the rodseal cavity of a test rig according to one embodiment;

FIGS. 41A and 41B show views of a rolling diaphragm in accordance withone embodiment;

FIGS. 42A and 42B show views of a rolling diaphragm in accordance withanother embodiment;

FIG. 43 shows a view of a double bellows system in accordance with oneembodiment;

FIGS. 44A and 44B show views of an airlock pressure regulation system inaccordance with one embodiment;

FIG. 45 shows a bidirectional regulator according to one embodiment;

FIGS. 46A-46E show various positions of a spool valve in a bidirectionalregulator in accordance with various embodiments;

FIG. 47 shows a view of an airlock pressure regulation system inaccordance with one embodiment;

FIG. 48 shows a view of an airlock pressure regulation system inaccordance with one embodiment;

FIG. 49 shows a view of a mechanical pump for regulating airlockpressure in accordance with one embodiment;

FIGS. 50A and 50B show views of a heat exchanger in accordance with oneembodiment;

FIGS. 51A and 51B show views of a rocking beam mechanism in accordancewith one embodiment;

FIGS. 52A and 52B show views of a horizontally supported Stirling cycleengine in accordance with one embodiment;

FIGS. 53A and 53B show views of a tube-in-tube heat exchanger accordingto one embodiment;

FIGS. 53C and 53D show views of a tube based heat exchanger according toone embodiment;

FIGS. 54-59 show various views of a burner in accordance with oneembodiment;

FIG. 60 show views of a burner in accordance with one embodiment;

FIG. 61 is a diagram of a control burner scheme in accordance with oneembodiment;

FIGS. 62A-62D are a further embodiment of a venturi-type burner for usein conjunction with the multiple heater head in accordance with oneembodiment;

FIG. 63 shows a further embodiment of an airlock pressure regulationsystem;

FIGS. 64A-64F present an embodiment of an Airlock delta PressureRegulation (AdPR) block;

FIG. 65A-65D is an embodiment of an Airlock delta Pressure Regulation(AdPR) block;

FIGS. 65E-65J is an embodiment of an Airlock delta Pressure Regulation(AdPR) block;

FIGS. 66A-66E show a further embodiment of various positions of a spoolvalve in a bidirectional regulator in accordance with variousembodiments;

FIG. 67A shows a chart which shows the order of the embodiments asdepicted in FIGS. 67B-H;

FIGS. 67 B-H illustrate an embodiment of a Stirling Engine Controller;

FIG. 68A is a cross-section of one embodiment of a Stirling Engine witha piston rod seal unit;

FIG. 68B is a detailed view of the piston rod seal unit in FIG. 68A;

FIG. 69A is a cross-section of one embodiments of a piston rod sealunit;

FIG. 69B is a cross-section of one embodiment of a floating rod sealassembly;

FIG. 69C is an isometric view of one embodiment of a piston rod sealunit;

FIG. 69C is an isometric view of one embodiment of a piston rod sealunit;

FIG. 69D is a cross-section of an embodiment of a piston rod seal unitwith clearance seals;

FIG. 69E is a cross-section of an embodiment of a piston rod seal unitwith clearance seals;

FIG. 69F is a cross-section of an embodiment of a piston rod seal unitwith a hybrid clearance and lip seal;

FIG. 70A is a cross-section of a further embodiment of aannular-venturi-type burner for use in conjunction with the multipleheater head in accordance with one embodiment;

FIG. 70B is a cross-section of a the annular-venturi-type burner head;

FIG. 70C is a detailed view of the annular-venturi with fuel ports andan ignitor;

FIG. 70D is a perspective view of a radial swirler at the entrance tothe annular-venturi;

FIG. 71A is a cross-section of one embodiment of a Stirling Engine withan linear-cross-head bearing;

FIG. 71B is a cross-section of one embodiment of a piston, piston rodand linear-cross-head bearing;

FIG. 71C is a cross-section of one embodiment of a linear-cross-headbearing; and

FIG. 72 is a cross-section of one embodiment of a piston with aclearance seal.

DETAILED DESCRIPTION

Stirling cycle machines, including engines and refrigerators, have along technological heritage, described in detail in Walker, StirlingEngines, Oxford University Press (1980), incorporated herein byreference. The principle underlying the Stirling cycle engine is themechanical realization of the Stirling thermodynamic cycle:isovolumetric heating of a gas within a cylinder, isothermal expansionof the gas (during which work is performed by driving a piston),isovolumetric cooling, and isothermal compression. Additional backgroundregarding aspects of Stirling cycle machines and improvements thereto isdiscussed in Hargreaves, The Phillips Stirling Engine (Elsevier,Amsterdam, 1991), which is herein incorporated by reference.

The principle of operation of a Stirling cycle machine is readilydescribed with reference to FIGS. 1A-1E, wherein identical numerals areused to identify the same or similar parts. Many mechanical layouts ofStirling cycle machines are known in the art, and the particularStirling cycle machine designated generally by numeral 10 is shownmerely for illustrative purposes. In FIGS. 1A to 1D, piston 12 and adisplacer 14 move in phased reciprocating motion within the cylinders 16which, in some embodiments of the Stirling cycle machine, may be asingle cylinder, but in other embodiments, may include greater than asingle cylinder. A working fluid contained within cylinders 16 isconstrained by seals from escaping around piston 12 and displacer 14.The working fluid is chosen for its thermodynamic properties, asdiscussed in the description below, and is typically helium at apressure of several atmospheres, however, any gas, including any inertgas, may be used, including, but not limited to, hydrogen, argon, neon,nitrogen, air and any mixtures thereof. The position of the displacer 14governs whether the working fluid is in contact with the hot interface18 or the cold interface 20, corresponding, respectively, to theinterfaces at which heat is supplied to and extracted from the workingfluid. The supply and extraction of heat is discussed in further detailbelow. The volume of working fluid governed by the position of thepiston 12 is referred to as the compression space 22.

During the first phase of the Stirling cycle, the starting condition ofwhich is depicted in FIG. 1A, the piston 12 compresses the fluid in thecompression space 22. The compression occurs at a substantially constanttemperature because heat is extracted from the fluid to the ambientenvironment. The condition of the Stirling cycle machine 10 aftercompression is depicted in FIG. 1B. During the second phase of thecycle, the displacer 14 moves in the direction of the cold interface 20,with the working fluid displaced from the region of the cold interface20 to the region of the hot interface 18. This phase may be referred toas the transfer phase. At the end of the transfer phase, the fluid is ata higher pressure since the working fluid has been heated at constantvolume. The increased pressure is depicted symbolically in FIG. 1C bythe reading of the pressure gauge 24.

During the third phase (the expansion stroke) of the Stirling cyclemachine, the volume of the compression space 22 increases as heat isdrawn in from outside the Stirling cycle machine 10, thereby convertingheat to work. In practice, heat is provided to the fluid by means of aheater head (not shown) which is discussed in greater detail in thedescription below. At the end of the expansion phase, the compressionspace 22 is full of cold fluid, as depicted in FIG. 1D. During thefourth phase of the Stirling cycle machine 10, fluid is transferred fromthe region of the hot interface 18 to the region of the cold interface20 by motion of the displacer 14 in the opposing sense. At the end ofthis second transfer phase, the fluid fills the compression space 22 andcold interface 20, as depicted in FIG. 1A, and is ready for a repetitionof the compression phase. The Stirling cycle is depicted in a P-V(pressure-volume) diagram as shown in FIG. 1E.

Additionally, on passing from the region of the hot interface 18 to theregion of the cold interface 20. In some embodiments, the fluid may passthrough a regenerator (shown as 408 in FIG. 4). A regenerator is amatrix of material having a large ratio of surface area to volume whichserves to absorb heat from the fluid when it enters from the region ofthe hot interface 18 and to heat the fluid when it passes from theregion of the cold interface 20.

Stirling cycle machines have not generally been used in practicalapplications due to several daunting challenges to their development.These involve practical considerations such as efficiency and lifetime.Accordingly, there is a need for more Stirling cycle machines withminimal side loads on pistons, increased efficiency and lifetime.

The principle of operation of a Stirling cycle machine or Stirlingengine is further discussed in detail in U.S. Pat. No. 6,381,958, issuedMay 7, 2002, to Kamen et al., which is herein incorporated by referencein its entirety.

Rocking Beam Drive

Referring now to FIGS. 2-4, embodiments of a Stirling cycle machine,according to one embodiment, are shown in cross-section. The engineembodiment is designated generally by numeral 300. While the Stirlingcycle machine will be described generally with reference to the Stirlingengine 300 embodiments shown in FIGS. 2-4, it is to be understood thatmany types of machines and engines, including but not limited torefrigerators and compressors may similarly benefit from variousembodiments and improvements which are described herein, including butnot limited to, external combustion engines and internal combustionengines.

FIG. 2 depicts a cross-section of an embodiment of a rocking beam drivemechanism 200 (the term “rocking beam drive” is used synonymously withthe term “rocking beam drive mechanism”) for an engine, such as aStirling engine, having linearly reciprocating pistons 202 and 204housed within cylinders 206 and 208, respectively. The cylinders includelinear bearings 220. Rocking beam drive 200 converts linear motions ofpistons 202 and 204 into the rotary motion of a crankshaft 214. Rockingbeam drive 200 has a rocking beam 216, rocker pivot 218, a firstcoupling assembly 210, and a second coupling assembly 212. Pistons 202and 204 are coupled to rocking beam drive 200, respectively, via firstcoupling assembly 210 and second coupling assembly 212. The rocking beamdrive is coupled to crankshaft 214 via a connecting rod 222.

In some embodiments, the rocking beam and a first portion of thecoupling assembly may be located in a crankcase, while the cylinders,pistons and a second portion of the coupling assembly is located in aworkspace.

In FIG. 4 a crankcase 400 most of the rocking beam drive 200 ispositioned below the cylinder housing 402. Crankcase 400 is a space topermit operation of rocking beam drive 200 having a crankshaft 214,rocking beam 216, linear bearings 220, a connecting rod 222, andcoupling assemblies 210 and 212. Crankcase 400 intersects cylinders 206and 208 transverse to the plane of the axes of pistons 202 and 204.Pistons 202 and 204 reciprocate in respective cylinders 206 and 208, asalso shown in FIG. 2. Cylinders 206 and 208 extend above crankshafthousing 400. Crankshaft 214 is mounted in crankcase 400 below cylinders206 and 208.

FIG. 2 shows one embodiment of rocking beam drive 200. Couplingassemblies 210 and 212 extend from pistons 202 and 204, respectively, toconnect pistons 202 and 204 to rocking beam 216. Coupling assembly 212for piston 204, in some embodiments, may comprise a piston rod 224 and alink rod 226. Coupling assembly 210 for piston 202, in some embodiments,may comprise a piston rod 228 and a link rod 230. Piston 204 operates inthe cylinder 208 vertically and is connected by the coupling assembly212 to the end pivot 232 of the rocking beam 216. The cylinder 208provides guidance for the longitudinal motion of piston 204. The pistonrod 224 of the coupling assembly 212 attached to the lower portion ofpiston 204 is driven axially by its link rod 226 in a substantiallylinear reciprocating path along the axis of the cylinder 208. The distalend of piston rod 224 and the proximate end of link rod 226, in someembodiments, may be jointly hinged via a coupling means 234. Thecoupling means 234, may be any coupling means known in the art,including but not limited to, a flexible joint, roller bearing element,hinge, journal bearing joint (shown as 600 in FIG. 6), and flexure(shown as 700 in FIGS. 7A and 7B). The distal end of the link rod 226may be coupled to one end pivot 232 of rocking beam 216, which ispositioned vertically and perpendicularly under the proximate end of thelink rod 226. A stationary linear bearing 220 may be positioned alongcoupling assembly 212 to further ensure substantially linearlongitudinal motion of the piston rod 224 and thus ensuringsubstantially linear longitudinal motion of the piston 204. In anexemplary embodiment, link rod 226 does not pass through linear bearing220. This ensures, among other things, that piston rod 224 retains asubstantially linear and longitudinal motion.

In the exemplary embodiment, the link rods may be made from aluminum,and the piston rods and connecting rod are made from D2 Tool Steel.Alternatively, the link rods, piston rods, connecting rods, and rockingbeam may be made from 4340 steel. Other materials may be used for thecomponents of the rocking beam drive, including, but not limited to,titanium, aluminum, steel or cast iron. In some embodiments, the fatiguestrength of the material being used is above the actual load experiencedby the components during operation.

Still referring to FIGS. 2-4, piston 202 operates vertically in thecylinder 206 and is connected by the coupling assembly 210 to the endpivot 236 of the rocking beam 216. The cylinder 206 serves, amongstother functions, to provide guidance for longitudinal motion of piston202. The piston rod 228 of the coupling assembly 210 is attached to thelower portion of piston 202 and is driven axially by its link rod 230 ina substantially linear reciprocating path along the axis of the cylinder206. The distal end of the piston rod 228 and the proximate end of thelink rod 230, in some embodiments, is jointly hinged via a couplingmeans 238. The coupling means 238, in various embodiments may include,but are not limited to, a flexure (shown as 700 in FIGS. 7A and 7B,roller bearing element, hinge, journal bearing (shown as 600 in FIG. 6),or coupling means as known in the art. The distal end of the link rod230, in some embodiments, may be coupled to one end pivot 236 of rockingbeam 216, which is positioned vertically and perpendicularly under theproximate end of link rod 230. A stationary linear bearing 220 may bepositioned along coupling assembly 210 to further ensure linearlongitudinal motion of the piston rod 228 and thus ensuring linearlongitudinal motion of the piston 202. In an exemplary embodiment, linkrod 230 does not pass through linear bearing 220 to ensure that pistonrod 228 retains a substantially linear and longitudinal motion.

The coupling assemblies 210 and 212 change the alternating longitudinalmotion of respective pistons 202 and 204 to oscillatory motion of therocking beam 216. The delivered oscillatory motion is changed to therotational motion of the crankshaft 214 by the connecting rod 222,wherein one end of the connecting rod 222 is rotatably coupled to aconnecting pivot 240 positioned between an end pivot 232 and a rockerpivot 218 in the rocking beam 216, and another end of the connecting rod222 is rotatably coupled to crankpin 246. The rocker pivot 218 may bepositioned substantially at the midpoint between the end pivots 232 and236 and oscillatorily support the rocking beam 216 as a fulcrum, thusguiding the respective piston rods 224 and 228 to make sufficient linearmotion. In the exemplary embodiment, the crankshaft 214 is located abovethe rocking beam 216, but in other embodiments, the crankshaft 214 maybe positioned below the rocking beam 216 (as shown in FIGS. 5B and 5D)or in some embodiments, the crankshaft 214 is positioned to the side ofthe rocking beam 216, such that it still has a parallel axis to therocking beam 216.

Still referring to FIGS. 2-4, the rocking beam oscillates about therocker pivot 218, the end pivots 232 and 236 follow an arc path. Sincethe distal ends of the link rods 226 and 230 are connected to therocking beam 216 at pivots 232 and 236, the distal ends of the link rods226 and 230 also follow this arc path, resulting in an angular deviation242 and 244 from the longitudinal axis of motion of their respectivepistons 202 and 204. The coupling means 234 and 238 are configured suchthat any angular deviation 244 and 242 from the link rods 226 and 230experienced by the piston rods 224 and 228 is minimized. Essentially,the angular deviation 244 and 242 is absorbed by the coupling means 234and 238 so that the piston rods 224 and 228 maintain substantiallylinear longitudinal motion to reduce side loads on the pistons 204 and202. A stationary linear bearing 220 may also be placed inside thecylinder 208 or 206, or along coupling assemblies 212 or 210, to furtherabsorb any angular deviation 244 or 242 thus keeping the piston push rod224 or 228 and the piston 204 or 202 in linear motion along thelongitudinal axis of the piston 204 or 202.

Therefore, in view of reciprocating motion of pistons 202 and 204, it isnecessary to keep the motion of pistons 202 and 204 as close to linearas possible because the deviation 242 and 244 from longitudinal axis ofreciprocating motion of pistons 202 and 204 causes noise, reduction ofefficiency, increase of friction to the wall of cylinder, increase ofside-load, and low durability of the parts. The alignment of thecylinders 206 and 208 and the arrangement of crankshaft 214, piston rods224 and 228, link rods 226 and 230, and connecting rod 222, hence, mayinfluence on, amongst other things, the efficiency and/or the volume ofthe device. For the purpose of increasing the linearity of the pistonmotion as mentioned, the pistons (shown as 202 and 204 in FIGS. 2-4) arepreferably as close to the side of the respective cylinders 206 and 208as possible.

In another embodiment reducing angular deviation of link rods, link rods226 and 230 substantially linearly reciprocate along longitudinal axisof motion of respective pistons 204 and 202 to decrease the angulardeviation and thus to decrease the side load applied to each piston 204and 202. The angular deviation defines the deviation of the link rod 226or 230 from the longitudinal axis of the piston 204 or 202. Numerals 244and 242 designate the angular deviation of the link rods 226 and 230, asshown in FIG. 2. Therefore, the position of coupling assembly 212influences the angular displacement of the link rod 226, based on thelength of the distance between the end pivot 232 and the rocker pivot218 of the rocking beam 216. Thus, the position of the couplingassemblies may be such that the angular displacement of the link rod 226is reduced. For the link rod 230, the length of the coupling assembly210 also may be determined and placed to reduce the angular displacementof the link rod 230, based on the length of the distance between the endpivot 236 and the rocker pivot 218 of the rocking beam 216. Therefore,the length of the link rods 226 and 230, the length of couplingassemblies 212 and 210, and the length of the rocking beam 216 aresignificant parameters that greatly influence and/or determine theangular deviation of the link rods 226 and 230 as shown in FIG. 2.

The exemplary embodiment has a straight rocking beam 216 having the endpoints 232 and 236, the rocker pivot 218, and the connecting pivot 240along the same axis. However, in other embodiments, the rocking beam 216may be bent, such that pistons may be placed at angles to each other, asshown in FIGS. 5C and 5D.

Referring now to FIGS. 2-4 and FIGS. 7A-7B, in some embodiments of thecoupling assembly, the coupling assemblies 212 and 210, may include aflexible link rod that is axially stiff but flexible in the rocking beam216 plane of motion between link rods 226 and 230, and pistons 204 and202, respectively. In this embodiment, at least one portion, the flexure(shown as 700 in FIGS. 7A and 7B), of link rods 226 and 230 is elastic.The flexure 700 acts as a coupling means between the piston rod and thelink rod. The flexure 700 may absorb the crank-induced side loads of thepistons more effectively, thus allowing its respective piston tomaintain linear longitudinal movement inside the piston's cylinder. Thisflexure 700 allows small rotations in the plane of the rocking beam 216between the link rods 226 and 230 and pistons 204 or 202, respectively.Although depicted in this embodiment as flat, which increases theelasticity of the link rods 226 and 230, the flexure 700, in someembodiments, is not flat. The flexure 700 also may be constructed nearto the lower portion of the pistons or near to the distal end of thelink rods 226 and 230. The flexure 700, in one embodiment, may be madeof #D2 Tool Steel Hardened to 58-62 RC. In some embodiments, there maybe more than one flexure (not shown) on the link rod 226 or 230 toincrease the elasticity of the link rods.

In alternate embodiment, the axes of the pistons in each cylinderhousing may extend in different directions, as depicted in FIGS. 5C and5D. In the exemplary embodiment, the axes of the pistons in eachcylinder housing are substantially parallel and preferably substantiallyvertical, as depicted in FIGS. 2-4, and FIGS. 5A and 5B. FIGS. 5A-5Dinclude various embodiments of the rocking beam drive mechanismincluding like numbers as those shown and described with respect toFIGS. 2-4. It will be understood by those skilled in that art thatchanging the relative position of the connecting pivot 240 along therocking beam 216 will change the stroke of the pistons.

Accordingly, a change in the parameters of the relative position of theconnecting pivot 240 in the rocking beam 216 and the length of thepiston rods 224 and 228, link rods 230 and 226, rocking beam 216, andthe position of rocker pivot 218 will change the angular deviation ofthe link rods 226 and 230, the phasing of the pistons 204 and 202, andthe size of the device 300 in a variety of manner. Therefore, in variousembodiments, a wide range of piston phase angles and variable sizes ofthe engine may be chosen based on the modification of one or more ofthese parameters. In practice, the link rods 224 and 228 of theexemplary embodiment have substantially lateral movement within from−0.5 degree to +0.5 degree from the longitudinal axis of the pistons 204and 202. In various other embodiments, depending on the length of thelink rod, the angle may vary anywhere from approaching 0 degrees to 0.75degrees. However, in other embodiments, the angle may be higherincluding anywhere from approaching 0 to the approximately 20 degrees.As the link rod length increases, however, the crankcase/overall engineheight increases as well as the weight of the engine.

One feature of the exemplary embodiment is that each piston has its linkrod extending substantially to the attached piston rod so that it isformed as a coupling assembly. In one embodiment, the coupling assembly212 for the piston 204 includes a piston rod 224, a link rod 226, and acoupling means 234 as shown in FIG. 2. More specifically, one proximalend of piston rod 224 is attached to the lower portion of piston 204 andthe distal end piston rod 224 is connected to the proximate end of thelink rod 226 by the coupling means 234. The distal end of the link rod226 extends vertically to the end pivot 232 of the rocking beam 216. Asdescribed above, the coupling means 234 may be, but is not limited to, ajoint, hinge, coupling, or flexure or other means known in the art. Inthis embodiment, the ratio of the piston rod 224 and the link rod 226may determine the angular deviation of the link rod 226 as mentionedabove.

Referring now to FIG. 4, one embodiment of the engine is shown. Here thepistons 202 and 204 of engine 300 operate between a hot chamber 404 anda cold chamber 406 of cylinders 206 and 208 respectively. Between thetwo chambers there may be a regenerator 408. The regenerator 408 mayhave variable density, variable area, and, in some embodiments, is madeof wire. The varying density and area of the regenerator may be adjustedsuch that the working gas has substantially uniform flow across theregenerator 408. Various embodiments of the regenerator 408 arediscussed in detail below, and in U.S. Pat. No. 6,591,609, issued Jul.17, 2003, to Kamen et al., and U.S. Pat. No. 6,862,883, issued Mar. 8,2005, to Kamen et al., which are herein incorporated by reference intheir entireties. When the working gas passes through the hot chamber404, a heater head 410 may heat the gas causing the gas to expand andpush pistons 202 and 204 towards the cold chamber 406, where the gascompresses. As the gas compresses in the cold chamber 406, pistons 202and 204 may be guided back to the hot chamber to undergo the Stirlingcycle again. The heater head 410 may have one of several forms includinga pin head, a fin head, a folded fin head, or heater tubes as shown inFIG. 4 or any other heater head embodiment known, including, but notlimited to, those described below. Various embodiments of heater head410 are discussed in detail below, and in U.S. Pat. No. 6,381,958,issued May 7, 2002, to Kamen et al., U.S. Pat. No. 6,543,215, issuedApr. 8, 2003, to Langenfeld et al., U.S. Pat. No. 6,966,182, issued Nov.22, 2005, to Kamen et al, and U.S. Pat. No. 7,308,787, issued Dec. 18,2007, to LaRocque et al., and in U.S. patent application Ser. No.13/447,990, filed Apr. 16, 2012 and entitled Stirling Cycle Machine(Attorney Docket No. 184), all of which are hereby incorporated hereinby reference in their entireties.

In some embodiments, a cooler 412 may be positioned alongside cylinders206 and 208 to further cool the gas passing through to the cold chamber406. Various embodiments of cooler 412 are discussed in detail in theproceeding sections, and in U.S. Pat. No. 7,325,399, issued Feb. 5,2008, to Strimling et al, which is herein incorporated by reference inits entirety.

In some embodiments, at least one piston seal 414 may be positioned onpistons 202 and 204 to seal the hot section 404 off from the coldsection 406. Additionally, at least one piston guide ring 416 may bepositioned on pistons 202 and 204 to help guide the pistons' motion intheir respective cylinders. Various embodiments of piston seal 414 andguide ring 416 are described in detail below, and in U.S. patentapplication Ser. No. 10/175,502, filed Jun. 19, 2002, published Feb. 6,2003 (now abandoned), which is herein incorporated by reference in itsentirety.

In some embodiments, at least one piston rod seal 418 may be placedagainst piston rods 224 and 228 to prevent working gas from escapinginto the crankcase 400, or alternatively into airlock space 420. Thepiston rod seal 418 may be an elastomer seal, or a spring-loaded seal.Various embodiments of the piston rod seal 418 are discussed in detailbelow.

In some embodiments, the airlock space may be eliminated, for example,in the rolling diaphragm and/or bellows embodiments described in moredetail below. In those cases, the piston rod seals 224 and 228 seal theworking space from the crankcase.

In some embodiments, at least one rolling diaphragm/bellows 422 may belocated along piston rods 224 and 228 to prevent airlock gas fromescaping into the crankcase 400. Various embodiments of rollingdiaphragm 422 are discussed in more detail below.

Although FIG. 4 shows a cross section of engine 300 depicting only twopistons and one rocking beam drive, it is to be understood that theprinciples of operation described herein may apply to a four cylinder,double rocking beam drive engine, as designated generally by numeral 800in FIG. 10B.

Piston Operation

Referring now to FIG. 8 that shows the operation of pistons 802, 804,806, and 808 during one revolution of crankshaft 814. With a ¼revolutionof crankshaft 814, piston 802 is at the top of its cylinder, otherwiseknown as top dead center, piston 806 is in upward midstroke, piston 804is at the bottom of its cylinder, otherwise known as bottom dead center,and piston 808 is in downward midstroke. With a ½ revolution ofcrankshaft 814, piston 802 is in downward midstroke, piston 806 is attop dead center, piston 804 is in upward midstroke, and piston 808 is atbottom dead center. With ¾revolution of crankshaft 814, piston 802 is atbottom dead center, piston 806 is in downward midstroke, piston 804 isat top dead center, and piston 808 is in upward midstroke. Finally, witha full revolution of crankshaft 814, piston 802 is in upward midstroke,piston 806 is at bottom dead center, piston 804 is in downwardmidstroke, and piston 808 is at top dead center. During each ¼revolution, there is a 90 degree phase difference between pistons 802and 806, a 180 degree phase difference between pistons 802 and 804, anda 270 degree phase difference between pistons 802 and 808. FIG. 9Aillustrates the relationship of the pistons being approximately 90degrees out of phase with the preceding and succeeding piston.Additionally, FIG. 8 shows the exemplary embodiment machine means oftransferring work. Thus, work is transferred from piston 802 to piston806 to piston 804 to piston 808 so that with a full revolution ofcrankshaft 814, all pistons have exerted work by moving from the top tothe bottom of their respective cylinders.

Referring now to FIG. 8, together with FIGS. 9A-9C, illustrate the 90degree phase difference between the pistons in the exemplary embodiment.Referring now to FIG. 9A, although the cylinders are shown in a linearpath, this is for illustration purposes only. In the exemplaryembodiment of a four cylinder Stirling cycle machine, the flow path ofthe working gas contained within the cylinder working space follows afigure eight pattern. Thus, the working spaces of cylinders 1200, 1202,1204, and 1206 are connected in a figure eight pattern, for example,from cylinder 1200 to cylinder 1202 to cylinder 1204 to cylinder 1208,the fluid flow pattern follows a figure eight. Still referring to FIG.9A, an unwrapped view of cylinders 1200, 1202, 1204, and 1206, takenalong the line B-B (shown in FIG. 9C) is illustrated. The 90 degreephase difference between pistons as described above allows for theworking gas in the warm section 1212 of cylinder 1204 to be delivered tothe cold section 1222 of cylinder 1206. As piston 802 and 808 are 90degrees out of phase, the working gas in the warm section 1214 ofcylinder 1206 is delivered to the cold section 1216 of cylinder 1200. Aspiston 802 and piston 806 are also 90 degrees out of phase, the workinggas in the warm section 1208 of cylinder 1200 is delivered to the coldsection 1218 of cylinder 1202. And as piston 804 and piston 806 are also90 degrees out of phase, so the working gas in the warm section 1210 ofcylinder 1202 is delivered to the cold section 1220 of cylinder 1204.Once the working gas of a warm section of a first cylinder enters thecold section of a second cylinder, the working gas begins to compress,and the piston within the second cylinder, in its down stroke,thereafter forces the compressed working gas back through a regenerator1224 and heater head 1226 (shown in FIG. 9B), and back into the warmsection of the first cylinder. Once inside the warm section of the firstcylinder, the gas expands and drives the piston within that cylinderdownward, thus causing the working gas within the cold section of thatfirst cylinder to be driven through the preceding regenerator and heaterhead, and into the cylinder. This cyclic transmigration characteristicof working gas between cylinders 1200, 1202, 1204, and 1206 is possiblebecause pistons 802, 804, 806, and 808 are connected, via drives 810 and812, to a common crankshaft 814 (shown in FIG. 8), in such a way thatthe cyclical movement of each piston is approximately 90 degrees inadvance of the movement of the proceeding piston, as depicted in FIG.9A.

Rolling Diaphragm, Metal Bellows, Airlock, and Pressure Regulator

In some embodiments of the Stirling cycle machine, lubricating fluid isused. To prevent the lubricating fluid from escaping the crankcase, aseal is used.

Referring now to FIGS. 10A-13B, some embodiments of the Stirling cyclemachine include a fluid lubricated rocking beam drive that utilizes arolling diaphragm 1300 positioned along the piston rod 1302 to preventlubricating fluid from escaping the crankcase, not shown, but thecomponents that are housed in the crankcase are represented as 1304, andentering areas of the engine that may be damaged by the lubricatingfluid. It is beneficial to contain the lubricating fluid for iflubricating fluid enters the working space, not shown, but thecomponents that are housed in the working space are represented as 1306,it would contaminate the working fluid, come into contact with theregenerator 1308, and may clog the regenerator 1308. The rollingdiaphragm 1300 may be made of an elastomer material, such as rubber orrubber reinforced with woven fabric or non-woven fabric to providerigidity. The rolling diaphragm 1300 may alternatively be made of othermaterials, such as fluorosilicone or nitrile with woven fabric ornon-woven fabric. The rolling diaphragm 1300 may also be made of carbonnanotubes or chopped fabric, which is non-woven fabric with fibers ofpolyester or KEVLAR®, for example, dispersed in an elastomer. In thesome embodiments, the rolling diaphragm 1300 is supported by the topseal piston 1328 and the bottom seal piston 1310. In other embodiments,the rolling diaphragm 1300 as shown in FIG. 10A is supported via notchesin the top seal piston 1328.

In some embodiments, a pressure differential is placed across therolling diaphragm 1300 such that the pressure above the seal 1300 isdifferent from the pressure in the crankcase 1304. This pressuredifferential inflates seal 1300 and allows seal 1300 to act as a dynamicseal as the pressure differential ensures that rolling diaphragmmaintains its form throughout operation. FIGS. 10A, and FIGS. 10C-10Dillustrate how the pressure differential effects the rolling diaphragm.The pressure differential causes the rolling diaphragm 1300 to conformto the shape of the bottom seal piston 1310 as it moves with the pistonrod 1302, and prevents separation of the seal 1300 from a surface of thepiston 1310 during operation. Such separation may cause seal failure.The pressure differential causes the rolling diaphragm 1300 to maintainconstant contact with the bottom seal piston 1310 as it moves with thepiston rod 1302. This occurs because one side of the seal 1300 willalways have pressure exerted on it thereby inflating the seal 1300 toconform to the surface of the bottom seal piston 1310. In someembodiments, the top seal piston 1328 ‘rolls over’ the corners of therolling diaphragm 1300 that are in contact with the bottom seal piston1310, so as to further maintain the seal 1300 in contact with the bottomseal piston 1310. In the exemplary embodiment, the pressure differentialis in the range of 10 to 15 PSI. The smaller pressure in the pressuredifferential is preferably in crankcase 1304, so that the rollingdiaphragm 1300 may be inflated into the crankcase 1304. However, inother embodiments, the pressure differential may have a greater orsmaller range of value.

The pressure differential may be created by various methods including,but not limited to, the use of the following: a pressurized lubricationsystem, a pneumatic pump, sensors, an electric pump, by oscillating therocking beam to create a pressure rise in the crankcase 1304, bycreating an electrostatic charge on the rolling diaphragm 1300, or othersimilar methods. In some embodiments, the pressure differential iscreated by pressurizing the crankcase 1304 to a pressure that is belowthe mean pressure of the working space 1306. In some embodiments thecrankcase 1304 is pressurized to a pressure in the range of 10 to 15 PSIbelow the mean pressure of the working space 1306, however, in variousother embodiments, the pressure differential may be smaller or greater.Further detail regarding the rolling diaphragm is included below.

Referring now to FIGS. 10C, and 11, however, another embodiment of theStirling machine is shown, wherein airlock space 1312 is located betweenworking space 1306 and crankcase 1304. Airlock space 1312 maintains aconstant volume and pressure necessary to create the pressuredifferential necessary for the function of rolling diaphragm 1300 asdescribed above. In one embodiment, airlock 1312 is not absolutelysealed off from working space 1306, so the pressure of airlock 1312 isequal to the mean pressure of working space 1306. Thus, in someembodiments, the lack of an effective seal between the working space andthe crankcase contributes to the need for an airlock space. Thus, theairlock space, in some embodiments, may be eliminated by a moreefficient and effective seal.

During operation, the working space 1306 mean pressure may vary so as tocause airlock 1312 mean pressure to vary as well. One reason thepressure may tend to vary is that during operation the working space mayget hotter, which in turn may increase the pressure in the workingspace, and consequently in the airlock as well since the airlock andworking space are in fluid communication. In such a case, the pressuredifferential between airlock 1312 and crankcase 1304 will also vary,thereby causing unnecessary stresses in rolling diaphragms 1300 that maylead to seal failure. Therefore, some embodiments of the machine, themean pressure within airlock 1312 is regulated so as to maintain aconstant desired pressure differential between airlock 1312 andcrankcase 1304, and ensuring that rolling diaphragms 1300 stay inflatedand maintains their form. In some embodiments, a pressure transducer isused to monitor and manage the pressure differential between the airlockand the crankcase, and regulate the pressure accordingly so as tomaintain a constant pressure differential between the airlock and thecrankcase. Various embodiments of the pressure regulator that may beused are described in further detail below, and in U.S. Pat. No.7,310,945, issued Dec. 25, 2007, to Gurski et al., which is hereinincorporated by reference in its entirety.

A constant pressure differential between the airlock 1312 and crankcase1304 may be achieved by adding or removing working fluid from airlock1312 via a pump or a release valve. Alternatively, a constant pressuredifferential between airlock 1312 and crankcase 1304 may be achieved byadding or removing working fluid from crankcase 1304 via a pump or arelease valve. The pump and release valve may be controlled by thepressure regulator. Working fluid may be added to airlock 1312 (orcrankcase 1304) from a separate source, such as a working fluidcontainer, or may be transferred over from crankcase 1304. Shouldworking fluid be transferred from crankcase 1304 to airlock 1312, it maybe desirable to filter the working fluid before passing it into airlock1312 so as to prevent any lubricant from passing from crankcase 1304into airlock 1312, and ultimately into working space 1306, as this mayresult in engine failure.

In some embodiments of the machine, crankcase 1304 may be charged with afluid having different thermal properties than the working fluid. Forexample, where the working gas is helium or hydrogen, the crankcase maybe charged with argon. Thus, the crankcase is pressurized. In someembodiments, helium is used, but in other embodiments, any inert gas, asdescribed herein, may be used. Thus, the crankcase is a wet pressurizedcrankcase in the exemplary embodiment. In other embodiments where alubricating fluid is not used, the crankcase is not wet.

In the exemplary embodiments, rolling diaphragms 1300 do not allow gasor liquid to pass through them, which allows working space 1306 toremain dry and crankcase 1304 to be wet sumped with a lubricating fluid.Allowing a wet sump crankcase 1304 increases the efficiency and life ofthe engine as there is less friction in rocking beam drives 1316. Insome embodiments, the use of roller bearings or ball bearings in drives1316 may also be eliminated with the use of lubricating fluid androlling diaphragms 1300. This may further reduce engine noise andincrease engine life and efficiency.

FIGS. 12A-12E show cross sections of various embodiments of the rollingdiaphragm (shown as 1400, 1410, 1412, 1422 and 1424) configured to bemounted between top seal piston and bottom seal piston (shown as 1328and 1310 in FIG. 10A), and between a top mounting surface and a bottommounting surface (shown as 1320 and 1318 in FIG. 10A). In someembodiments, the top mounting surface may be the surface of an airlockor working space, and the bottom mounting surface may be the surface ofa crankcase.

FIG. 12A shows one embodiment of the rolling diaphragm 1400, where therolling diaphragm 1400 includes a flat inner end 1402 that may bepositioned between a top seal piston and a bottom seal piston, so as toform a seal between the top seal piston and the bottom seal piston. Therolling diaphragm 1400 also includes a flat outer end 1404 that may bepositioned between a top mounting surface and a bottom mounting surface,so as to form a seal between the top mounting surface and the bottommounting surface. FIG. 12B shows another embodiment of the rollingdiaphragm, wherein rolling diaphragm 1410 may include a plurality ofbends 1408 leading up to flat inner end 1406 to provide for additionalsupport and sealing contact between the top seal piston and the bottomseal piston. FIG. 12C shows another embodiment of the rolling diaphragm,wherein rolling diaphragm 1412 includes a plurality of bends 1416leading up to flat outer end 1414 to provide for additional support andsealing contact between the top mounting surface and the bottom mountingsurface.

FIG. 12D shows another embodiment of the rolling diaphragm where rollingdiaphragm 1422 includes a bead along an inner end 1420 thereof, so as toform an ‘o-ring’ type seal between a top seal piston and a bottom sealpiston, and a bead along an outer end 1418 thereof, so as to form an‘o-ring’ type seal between a bottom mounting surface and a top mountingsurface. FIG. 12E shows another embodiment of the rolling diaphragm,wherein rolling diaphragm 1424 includes a plurality of bends 1428leading up to beaded inner end 1426 to provide for additional supportand sealing contact between the top seal piston and the bottom sealpiston. Rolling diaphragm 1424 may also include a plurality of bends1430 leading up to beaded outer end 1432 to provide for additionalsupport and sealing contact between the top seal piston and the bottomseal piston.

Although FIGS. 12A through 12E depict various embodiments of the rollingdiaphragm, it is to be understood that rolling diaphragms may be held inplace by any other mechanical means known in the art. Alternatively, therolling diaphragm may be replaced by metal bellows as disclosed in U.S.patent application Ser. No. 13/447,990, filed Apr. 16, 2012 and entitledStirling Cycle Machine (Attorney Docket No. 184), which is herebyincorporated herein by reference in its entirety.

Rolling Diaphragm and/or Bellows Embodiments

Various embodiments of the rolling diaphragm and/or bellows, whichfunction to seal, are described above. Further embodiments will beapparent to those of skill in the art based on the description above andthe additional description below relating to the parameters of therolling diaphragm and/or bellows.

In some embodiments, the pressure atop the rolling diaphragm or bellows,in the airlock space or airlock area (both terms are usedinterchangeably), is the mean-working-gas pressure for the machine,which, in some embodiments is an engine, while the pressure below therolling diaphragm and/or bellows, in the crankcase area, isambient/atmospheric pressure. In these embodiments, the rollingdiaphragm and/or bellows is required to operate with as much as 3000 psiacross it (and in some embodiments, up to 1500 psi or higher). In thiscase, the rolling diaphragm and/or bellows seal forms the working gas(helium, hydrogen, or otherwise) containment barrier for the machine(engine in the exemplary embodiment). Also, in these embodiments, theneed for a heavy, pressure-rated, structural vessel to contain thebottom end of the engine is eliminated, since it is now required tosimply contain lubricating fluid (oil is used as a lubricating fluid inthe exemplary embodiment) and air at ambient pressure, like aconventional internal combustion (“IC”) engine.

The capability to use a rolling diaphragm and/or bellows seal with suchan extreme pressure across it depends on the interaction of severalparameters. Referring now to FIG. 13A, an illustration of the actualload on the rolling diaphragm or bellows material is shown. As shown,the load is a function of the pressure differential and the annular gaparea for the installed rolling diaphragm or bellows seal.

Region 1 represents the portions of the rolling diaphragm and/or bellowsthat are in contact with the walls formed by the piston and cylinder.The load is essentially a tensile load in the axial direction, due tothe pressure differential across the rolling diaphragm and/or bellows.This tensile load due to the pressure across the rolling diaphragmand/or bellows can be expressed as:

L _(t) =P _(d) *A _(a)

Where

-   -   L_(t)=Tensile Load and    -   P_(d)=Pressure Differential    -   A_(a)=Annular Area

and

A _(a) =p/4*(D ² −d ²)

Where

-   -   D=Cylinder Bore and    -   d=Piston Diameter

The tensile component of stress in the bellows material can beapproximated as:

S _(t) =L _(t)/(p*(D+d)*t _(b))

Which reduces to:

S _(t) =P _(d)/4*(D−d)/tb

Later, we will show the relationship of radius of convolution, R_(c), toCylinder bore (D) and Piston Diameter (d) to be defined as:

R _(c)=(D−d)/4

So, this formula for St reduces to its final form:

S _(t) =P _(d) *R _(c) /t _(b)

Where

-   -   t_(b)=thickness of bellows material

Still referring to FIG. 13A, Region 2 represents the convolution. As therolling diaphragm and/or bellows material turns the corner, in theconvolution, the hoop stress imposed on the rolling diaphragm and/orbellows material may be calculated. For the section of the bellowsforming the convolution, the hoop component of stress can be closelyapproximated as:

S _(h) =P _(d) *R _(c) /t _(b)

The annular gap that the rolling diaphragm and/or bellows rolls withinis generally referred to as the convolution area. The rolling diaphragmand/or bellows fatigue life is generally limited by the combined stressfrom both the tensile (and hoop) load, due to pressure differential, aswell as the fatigue due to the bending as the fabric rolls through theconvolution. The radius that the fabric takes on during this ‘rolling’is defined here as the radius of convolution, Rc.

R _(c)=(D−d)/4

The bending stress, Sb, in the rolling diaphragm and/or bellows materialas it rolls through the radius of convolution, Rc, is a function of thatradius, as well as the thickness of the materials in bending. For afiber-reinforced material, the stress in the fibers themselves (duringthe prescribed deflection in the exemplary embodiments) is reduced asthe fiber diameter decreases. The lower resultant stress for the samelevel of bending allows for an increased fatigue life limit. As thefiber diameter is further reduced, flexibility to decrease the radius ofconvolution Rc is achieved, while keeping the bending stress in thefiber under its endurance limit. At the same time, as Rc decreases, thetensile load on the fabric is reduced since there is less unsupportedarea in the annulus between the piston and cylinder. The smaller thefiber diameter, the smaller the minimum Rc, the smaller the annulararea, which results in a higher allowable pressure differential.

For bending around a prescribed radius, the bending moment isapproximated by:

M=E*I/R

Where:

-   -   M=Bending Moment    -   E=Elastic Modulus    -   I=Moment of Inertia    -   R=Radius of Bend

Classical bending stress, S_(b), is calculated as:

S _(b) =M*Y/I

Where:

-   -   Y=Distance above neutral axis of bending

Substituting yields:

S _(b)=(E*I/R)*Y/I

S _(b) =E*Y/R

Assuming bending is about a central neutral axis:

Y _(max) =t _(b)/2

S _(b) =E*t _(b)/(2*R)

In some embodiments, rolling diaphragm and/or bellows designs for highcycle life are based on geometry where the bending stress imposed iskept about one order of magnitude less than the pressure-based loading(hoop and axial stresses). Based on the equation: Sb=E*tb/(2*R), it isclear that minimizing tb in direct proportion to Rc should not increasethe bending stress. The minimum thickness for the exemplary embodimentsof the rolling diaphragm and/or bellows material or membrane is directlyrelated to the minimum fiber diameter that is used in the reinforcementof the elastomer. The smaller the fibers used, the smaller resultant Rcfor a given stress level.

Another limiting component of load on the rolling diaphragm and/orbellows is the hoop stress in the convolution (which is theoreticallythe same in magnitude as the axial load while supported by the piston orcylinder). The governing equation for that load is as follows:

Sh=Pd*Rc/tb

Thus, if Rc is decreased in direct proportion to tb, then there is noincrease of stress on the membrane in this region. However, if thisratio is reduced in a manner that decreases Rc to a greater ratio thantb then parameters must be balanced. Thus, decreasing tb with respect toRc requires the rolling diaphragm and/or bellows to carry a heavierstress due to pressure, but makes for a reduced stress level due tobending. The pressure-based load is essentially constant, so this may befavorable—since the bending load is cyclic, therefore it is the bendingload component that ultimately limits fatigue life.

For bending stress reduction, tb ideally should be at a minimum, and Rcideally should be at a maximum. E ideally is also at a minimum. For hoopstress reduction, Rc ideally is small, and tb ideally is large.

Thus, the critical parameters for the rolling diaphragm and/or bellowsmembrane material are:

E, Elastic Modulus of the membrane material;

tb, membrane thickness (and/or fiber diameter);

Sut, Ultimate tensile strength of the rolling diaphragm and/or bellows;and

Slcf, The limiting fatigue strength of the rolling diaphragm and/orbellows.

Thus, from E, tb and Sut, the minimum acceptable Rc may be calculated.Next, using Rc, Slcf, and tb, the maximum Pd may be calculates. Rc maybe adjusted to shift the bias of load (stress) components between thesteady state pressure stress and the cyclic bending stress. Thus, theideal rolling diaphragm and/or bellows material is extremely thin,extremely strong in tension, and very limber in flexion.

Thus, in some embodiments, the rolling diaphragm and/or bellows material(sometimes referred to as a “membrane”), is made from carbon fibernanotubes. However, additional small fiber materials may also be used,including, but not limited to nanotube fibers that have been braided,nanotube untwisted yarn fibers, or any other conventional materials,including but not limited to KEVLAR, glass, polyester, synthetic fibersand any other material or fiber having a desirable diameter and/or otherdesired parameters as described in detail above.

Piston Seals and Piston Rod Seals

Referring now to FIG. 11, an embodiment of the machine is shown whereinan engine 1326, such as a Stirling cycle engine, includes at least onepiston rod seal 1314, a piston seal 1324, and a piston guide ring 1322,(shown as 1616 in FIG. 14). Various embodiments of the piston seal 1324and the piston guide ring 1322 are further discussed below, and in U.S.patent application Ser. No. 10/175,502 (now abandoned), which, asmentioned before, is incorporated by reference.

FIG. 14 shows a partial cross section of the piston 1600, driven alongthe central axis 1602 of cylinder, or the cylinder 1604. The piston seal(shown as 1324 in FIG. 11) may include a seal ring 1606, which providesa seal against the contact surface 1608 of the cylinder 1604. Thecontact surface 1608 is typically a hardened metal (preferably 58-62 RC)with a surface finish of 12 RMS or smoother. The contact surface 1608may be metal which has been case hardened, such as 8260 hardened steel,which may be easily case hardened and may be ground and/or honed toachieve a desired finish. The piston seal may also include a backingring 1610, which is sprung to provide a thrust force against the sealring 1606 thereby providing sufficient contact pressure to ensuresealing around the entire outward surface of the seal ring 1606. Theseal ring 1606 and the backing ring 1610 may together be referred to asa piston seal composite ring. In some embodiments, the at least onepiston seal may seal off a warm portion of cylinder 1604 from a coldportion of cylinder 1604.

Referring now to FIG. 15, some embodiments include a piston rod seal(shown as 1314 in FIG. 11) mounted in the piston rod cylinder wall 1700,which, in some embodiments, may include a seal ring 1706, which providesa seal against the contact surface 1708 of the piston rod 1604 (shown as1302 in FIG. 11). The contact surface 1708 in some embodiments is ahardened metal (preferably 58-62 RC) with a surface finish of 12 RMS orsmoother. The contact surface 1708 may be metal which has been casehardened, such as 8260 hardened steel, which may be easily case hardenedand may be ground and/or honed to achieve a desired finish. The pistonseal may also include a backing ring 1710, which is sprung to provide aradial or hoop force against the seal ring 1706 thereby providingsufficient contact hoop stress to ensure sealing around the entireinward surface of seal ring 1706. The seal ring 1706 and the backingring 1710 may together be referred to as a piston rod seal compositering.

In some embodiments, the seal ring and the backing ring may bepositioned on a piston rod, with the backing exerting an outwardpressure on the seal ring, and the seal ring may come into contact witha piston rod cylinder wall 1702. These embodiments require a largerpiston rod cylinder length than the previous embodiment. This is becausethe contact surface on the piston rod cylinder wall 1702 will be longerthan in the previous embodiment, where the contact surface 1708 lies onthe piston rod itself. In yet another embodiment, piston rod seals maybe any functional seal known in the art including, but not limited to,an o-ring, a graphite clearance seal, graphite piston in a glasscylinder, or any air pot, or a spring energized lip seal. In someembodiments, anything having a close clearance may be used, in otherembodiments, anything having interference, for example, a seal, is used.In the exemplary embodiment, a spring energized lip seal is used. Anyspring energized lip seal may be used, including those made by BAL SEALEngineering, Inc., Foothill Ranch, Calif. In some embodiments, the sealused is a BAL SEAL Part Number X558604.

The material of the seal rings 1606 and 1706 is chosen by considering abalance between the coefficient of friction of the seal rings 1606 and1706 against the contact surfaces 1608 and 1708, respectively, and thewear on the seal rings 1606 and 1706 it engenders. In applications inwhich piston lubrication is not possible, such as at the high operatingtemperatures of a Stirling cycle engine, the use of engineering plasticrings is used. The embodiments of the composition include a nylon matrixloaded with a lubricating and wear-resistant material. Examples of suchlubricating materials include PTFE/silicone, PTFE, graphite, etc.Examples of wear-resistant materials include glass fibers and carbonfibers. Examples of such engineering plastics are manufactured by LNPEngineering Plastics, Inc. of Exton, Pa. Backing rings 1610 and 1710 ispreferably metal.

The fit between the seal rings 1606 and 1706 and the seal ring grooves1612 and 1712, respectively, is preferably a clearance fit (about0.002″), while the fit of the backing rings 1610 and 1710 is preferablya looser fit, of the order of about 0.005″ in some embodiments. The sealrings 1606 and 1706 provide a pressure seal against the contact surfaces1608 and 1708, respectively, and also one of the surfaces 1614 and 1714of the seal ring grooves 1612 and 1712, respectively, depending on thedirection of the pressure difference across the rings 1606 and 1706 andthe direction of the piston 1600 or the piston rod 1704 travel.

Referring again to FIG. 14, at least one guide ring 1616 may also beprovided, in accordance with some embodiments, for bearing any side loadon piston 1600 as it moves up and down the cylinder 1604. Guide ring1616 is also preferably fabricated from an engineering plastic materialloaded with a lubricating material. A perspective view of guide ring1616 is shown in FIGS. 16A, 16B. An overlapping joint 2100 is shown andmay be diagonal to the central axis of guide ring 1616. The guide ring1616 and other guide elements located on the moving piston or on mattingstationary parts can bear significant side loads that are larger thanthe side loads supported by the pressure seal sand the associatedbacking ring. In a preferred embodiment, the backing ring 1610 exertsless than 15 psi, when fully compressed. The backing ring 1610 andpressure seal 1606 can bear a significant side load and thus are notdesigned to guide the piston 1600.

Referring now to FIG. 68A, which shows a cross-section of an embodimentof a Stirling engine and burner, in some embodiments, the length of timeof performance, i.e., the life, of the rod seals on the piston rod maybe extended and/or maximized by reducing and/or minimizing the axialmisalignment between the rod seals and the piston rod. In someembodiments, a rod seal may float radially when the pressure differencebetween the workspace and air lock is low, while forming a gas seal whenthe pressure difference between the workspace and air lock is high. Arod seal assembly comprising a cylinder-gland housing between theworking space and the air lock space is configured to receive thereciprocating piston rod that is disposed within the workspace and theairlock space. A floating bushing is configured to move axially andradially within the cylinder gland and is disposed coaxially around thereciprocating rod. A rod seal is configured to seal the outside diameterof the reciprocating rod relative to an inside surface of the floatingbushing and at least one stationary bushing fixed within the housingthat may form a seal with the floating bushing to limit the flow of gasbetween the workspace and the airlock. The rod seal may be a springenergized seal. The floating bushing may further comprises acircumferential flange on the outside surface that is configured toextend into the annular space formed by two stationary bushings and forma seal with one of the stationary bushings in the presence of a pressuredifference between the workspace and the airlock.

The rod seal assembly may also comprise a scraper ring located coaxiallyaround the piston rod and located within the cylinder-gland housingbetween the floating seal and the workspace. The cylinder-gland housingmay include a port connecting the workspace to an annular gap around thereciprocating rod between the scraper ring and the floating seal, thatminimizes the pressure difference across the scraper ring. The rod sealassembly may also comprise a magnetic particle trap between the scraperring and the floating seal.

In another embodiment, the rod seal assembly comprising a cylinder-glandhousing between the working space and the air lock space is configuredto receive the reciprocating piston rod that is disposed within theworkspace and the airlock space. A floating clearance bushing configuredto move axially and radially within the housing and is disposedcoaxially around the reciprocating piston rod and forms a clearance sealwith the piston rod. The cylinder-gland contains at least one stationaryannular element that is fixed within the housing and configured to forma face seal with the floating clearance bushing.

Referring now to FIGS. 68A-69F, a piston rod seal unit 13750 is shownthat allows the rod seal 13770 to move radially in order to minimizeforces on and maximize the life of the rod seal 13770. In addition, thepiston rod seal unit 13750 includes components to protect the rod seals13770 from particles. In various embodiments, the piston rod seal units13750 mount in the duct plate 13715 of the Stirling engine drive andprovides gas seals between the working spaces under the pistons 13738from the airlock 13736 on each piston rod 13744. The piston rod sealunit 13750 is presented in FIGS. 69A-69C with the piston rod 13744removed for clarity, wherein the rod seal 13770 is mounted in a floatingrod seal assembly 13760 and one or more floating rod seal assemblies13760 are mounted in the cylinder gland 13752. The pressure in theworkspaces may vary ±300 psi about the airlock pressure which requiresthat the rod seals 13770 minimize gas leaks in both directions. A smallamount of leakage between the workspace 13738 and the airlock 13736 istolerable and even desirable to allow the average pressure of each ofthe multiple workspaces 13738 to equalize with the pressure in thesingle airlock 13736 and thereby with each other. A large gas leakacross any of the piston rod seal units 13750 will reduce the pressureswing in the workspace and thereby the power and efficiency of theStirling engine. The sliding motion between the piston rod 13744 and therod seal 13770 may result in wear and abrasion of the seal surface thatleads to leaks, seal failure and reduced engine power and efficiency.

Still referring to FIGS. 68A-69C, to minimize the wear, the rod seal13770, in some embodiments, may be advantageously located concentricwith the piston rod 13744 so that the active sealing surfaces of rodseal 13770 are uniformly pressed against piston rod 13744. In someembodiments, and as shown in FIG. 68A, the location and motion path ofthe piston rod 13744 may be constrained by the crosshead bearing 13746at one end and the piston guide ring 13742 in the cylinder at the otherend. Misalignment or radial movement of either the crosshead bearing13746 or the piston guide ring 13742 may result in the piston rod 13744not being centered in mounting structure for the rod seal 13770 (e.g.the cylinder gland 13752) or in slight movement of the piston rod 13744during the stroke. The floating rod seal assembly 13760 (FIG. 69B)allows the rod seal 13770 to move radially in order to center itself onthe piston rod 13744 or minimize the variation of radial forces on thesealing surfaces around the circumference of the rod seal 13770. Thefloating rod seal assembly 13760 may also maintain a significant seal,but in some embodiments, not a perfect seal, to the flow of working gasbetween the workspace 13738 and airlock 13736, while allowing the radialmovement of the piston rod seal 13370.

Still referring to FIGS. 68A-69C, the piston rod seal units 13750including one or more floating rod seal assemblies 13760 mounted in thecylinder gland 13752, may be bolted to the duct plate 13715. In someembodiments, the piston rod seal units 13750 may be located to alignwith the heater heads 13712 and or the engine block 13741 that guidesthe crosshead bearings 13746 via structural features on the cylindergland 13752 and structural elements that assure proper alignment of ductplate 13715 with the cooler plate 13717 and engine block 13741. In someembodiments, the diameter 13784 on the cylinder gland 13752 mates with acounter-bored diameter on the duct plate 13715.

Still referring to FIGS. 68A-69C, in some embodiments, the piston rodseal unit 13750 includes; a housing 13754, into which one or morefloating rod seal assemblies 13760 are pressed and axially constrainedtherein, a scraper ring 13778, a particle trap 13780 and a port 13782 tominimize the pressure difference across the scraper ring 13778. Thefloating rod seal assemblies 13760 include an outer ring 13762, at leastone bushing 13764, (wherein, in some embodiments, may include twobushings 13764), a floating bushing 13766 and the rod seal 13770. Insome embodiments, the elements of the floating rod seal assembly 13760may be assembled by pressing a first bushing 13764 into the outer ring13762 until the one end of the bushing 13764 is flush with one end ofthe outer ring 13762; placing the floating bushing 13766 in the outerring 13762 on the non-flush end of the first bushing 13764; pressing thesecond bushing 13764 into the outer ring 13762 unit one end of thesecond bushing 13764 is flush with one end of the outer ring 13762 andthe floating bushing 13766 is captured between the two bushings 13764;and pressing the rod seal 13770 into the floating bushing 13766

Still referring to FIGS. 68A-69C, in various embodiments, the elementsof the floating rod seal assembly 13760 may be sized to allow thefloating bushing 13766 to move axially and radially, while rigidlyholding the bushings 13764 relative to the outer ring 13762. In variousembodiments, the outside diameter of the bushings 13764 and innerdiameter of the outer ring 13762 may be sized to provide a lightinterference fit, so that the bushings do not move relative outer ring13762. In some embodiments, the outside diameter of the bushings 13764and inner diameter of the outer ring 13762 may be sized to provide alocation fit. In various embodiments, the heights of the outer ring13762 and the two bushings 13764 and the axial thickness of the rib13766A may be sized so that the axial gap between the assembled bushings13764 is larger than the axial thickness of the rib 13766A. Theresulting axial gap 13768 may range from 0.001 to 0.02 inches. In someembodiments, the axial gap 13768 ranges from about 0.002 to about 0.004inches. In various embodiments, the outside diameter of the outer ring13762 and the inner diameter of the housing 13754 are selected toprovide leak tight seal. In various embodiments, the outside diameter ofthe outer ring 13762 and the inner diameter of housing 13754 may beselected to have a light interference fit. In another embodiment, theoutside diameter of the outer ring 13762 and the inner diameter ofhousing 13754 may be selected to have a location fit.

Still referring to FIGS. 68A-69C, in various embodiments, the outer ring13762 and housing 13754 may be made from metal. In various embodiments,the bushings 13764, 13766 may be made from metal with high wearresistance. In various embodiments, the cylinder gland 13752 may be madeof metal or high strength plastics that show resistance to cracking andfatigue failure. In some embodiments, the cylinder gland 13752 may be4140 steel. In various embodiments, the bushings 13764, 13766, may bemade of a wear resistant steel such as P20 (i.e. vacuum treated 4140steel) or a metal with similar strength and wear resistances. In variousembodiments, the outer ring 13762 and housing may be made from 4140 or4340 steel that has been treated to achieve a Rockwell Hardness of 28-32or a metal with similar strength and hardness. The surfaces on thefloating bushing 13766 and the bushings 13764 that contact each other(e.g. 13769 in FIG. 69B) provide a metal to metal seal. In variousembodiments, the contacting surfaces of the floating bushing 13766 andbushings 13764 may have a smooth surface finish. In some embodiments,the surface finish may be less than 32 microinches RA. In anotherexample the surface finish is less than 16 microinches RA.

Still referring to FIGS. 68A-69C, in various embodiments, the rod seal13770 may be a spring energized seal. In some embodiments, coiledsprings 13770A wrapped around the bearing surface may urge the lips ofthe seal 13770B toward the piston rod 13744. In some embodiments, thebearing surface may be composed of a composite polymer. In someembodiments, the bearing material may be a PFTE composite. In someembodiments, the rod seal 13770 may include an o-ring 13770C around theoutside diameter to provide a seal to axial flow of gas on the outsidediameter of the rod seal 13370. In some embodiments, the seal may begraphite or a graphite impregnated with antimony and or may be made ofany material and/or any combination and/or composite of material. Insome embodiments, the seal may be made of antimony impregnated graphitewhich may be from the SGL Carbon Company of Germany. In someembodiments, the rod seal may be a part supplied by CoorsTek of GoldenColo. or Bal Seal of Foothill Ranch, Calif.

Still referring to FIGS. 68A-69C in some embodiments, the floating rodseal assemblies may be assembled and axially constrained in the housing13754 following a method of: pressing the first floating rod sealassembly 13760 into the housing 13754 from the flanged end 13754A untilit is flush with the flanged end 13754A; pressing the washer 13758 intothe housing 13754 from the flanged end 13754A until it is flush with theflanged end 13754A; placing a second washer 13758 on other end of thefirst floating rod seal assembly 13760; pressing the second floating rodseal assembly 13760 into the housing 13754 from threaded end 13754Bplace; placing the third washer 13758 on the second floating rod sealassembly 13760; and threading the fitting 13776 into the housing 13754to axially secure the assemblies and washer 13760, 13758. In someembodiments, the housing 13754 may be secured to a rigid surface whilethe fitting 13776 is threaded into the housing 13754 and tightened to apredetermined torque. In various embodiments, the floating road sealassemblies 13760, washers 13758 and fitting 13776 may be assembled inusing various methods that would be evident to one skilled in the art.

Still referring to FIGS. 68A-69C in some embodiments, the piston rodseal unit 13750 may be assembled by pressing or inserting some or all ofthe following into the cylinder gland 13752: housing 13754 with thefloating rod seal assemblies 13760; and/or the particle trap 13780;and/or a scraper ring 13778. In various embodiments, an end plate 13756may be bolted to the bottom of the cylinder gland 13752 to capture thefloating rod seal assembles 13760, particle trap 13780 and/or scraperring 13778. In various embodiments, the port 13782 may be threaded intothe side of the cylinder gland between the scraper ring 13778 and thehousing 13754 containing the floating seals, so that both sides of thescraper ring 13778 are exposed to the same work space pressure. The port13782 includes a port 13782A and a particle filter 13782B that, in someembodiments, are separate pieces. In various embodiments, one or morecircumferential O-rings may be located between the housing 13754 and thecylinder gland 13752, and between the housing 13754 and the fitting13776 to seal leak paths through piston rod seal unit 137750. In someembodiments, there may be axial or face O-rings between the end plate13756 and the housing 13754 and between the end plate 13756 and thefirst washer 13758 to seal leak paths through piston rod seal unit137750.

In various embodiments, the floating rod seal assembly 13760 may beassembled differently such as pressing the elements into the housing13754. In another embodiment the outer ring 13762 may be incorporatedinto the housing 13754, so the bushings 13764, floating bushing 13766and rod seal 13770 may be directly mounted into the housing 13754. Inanother embodiment, the floating rod seal assemblies 13760 may bepressed or inserted directly into the cylinder gland 13752 or the ductplate 13715.

Still referring to FIGS. 68A-69C, the scraper ring 13778, particle trap13780, port 13782, and filter 13782B, may serve to protect the rod seal13370 and floating rod seal assemblies from particles and/or may preventparticles from entering the piston rod seal unit 13750 with the movementof gas into the piston rod seal unit 13750. Particles may be produced inthe workspace, for example, in some embodiments when pieces of the finewire break off the regenerator. In some embodiments metal particles maybe generated when the piston contacts the cylinder. In some embodiments,metal flakes may be generated from oxidation of metal in the heater heador other causes. Metal and/or oxidized metal particles that reach therod seal 13370 may score the piston rod 13744 or damage the rod seal13770 leading to leakage and lower power or early seal failure.Particles may also get trapped in the axial gap 13768 between thebushing and the floating bushing 13766, which may create a leak patharound the rod seal 13770 that may reduce engine power. Therefore, invarious embodiments, it is desirable to minimize and/or prevent metalparticles from entering piston rod seal space. In various embodiments,the scraper ring 13778 and/or the filter 13782B may minimize the numberof particles that enter the piston rod seal unit 13750. The particletrap 13780 may serve to attract and hold particles that pass by thescraper ring 13778 or filter 13782B. In some embodiments, the particletrap may include one or more magnets 13780A that may attract steelparticles and oxides of metal, which may include iron.

In various embodiments, the port 13782 may prevent and/or minimizeand/or reduce a pressure difference across the scraper ring 13778 byfluidly connecting the bottom side of the scraper ring 13778 to the sameworkspace 13738 in which the top of the scraper ring 13778 may beexposed. In various embodiments, the port may prevent and/or minimizeand/or reduce a pressure difference across the scraper ring 13778 due tochanges in the average workspace pressure which may occur in variousstages, for example, but not limited to, during startup and/or due tothe action of the scraper ring 13778 allowing flow, for example, in onedirection along the piston rod 13744, but not allowing flow in the otherdirection. Still referring to FIGS. 68A-69C, in some embodiments, thescraper ring 13778 may be a spring energized lip seal. In someembodiments, coiled springs wrapped around the bearing surface urge theends of the bearing surface toward the piston rod 13744 and the bearingsurface. In various embodiments, the bearing surface may be made of acomposite polymer. In some embodiments, the bearing surface may be madefrom a PFTE composite. In some embodiments, the rod seal 13370 may beone supplied by CoorsTek of Golden Colo. or Bal Seal Engineering Inc. ofFoothill Ranch, Calif.

Still referring to FIGS. 68A-69C, in some embodiments, the floating rodseal assembly 13760 allows the rod seals 13370 to self-center on thereciprocating piston rod 13744. The floating bushing 13766, which issealed to the outside diameter of the rod seal 13770 may form a faceseal against one of the stationary bearings 13764, 13765 to keep gasfrom leaking around the rod seal. In some embodiments, the axialmovement of the floating bushing 13766 may be constrained by the a smallaxial distance between the two stationary bushings 13764, 13765 that is0.002 to 0.004 larger than the thickness of the circumferential rib13766A of the floating bushing 13766.

Still referring to FIGS. 68A-69C, in some embodiments, the operation ofthe floating bushing 13766 may be best understood by considering itsmotion as the axial pressure difference changes direction during everystroke. Herein, the axial pressure difference is the difference inpressure along the axis of the piston rod 13744 from one side of the rodseal 13770 or floating bushing 13766 to the other. The sealing action ofthe floating rod seal assemblies 13760 may be understood by consideringone piston rod 13744 and the mated piston rod seal unit 13760, when thepiston rod 13744 is at either the top or the bottom of the its stroke.At this point, there exists a large axial pressure difference acrosseach of the floating rod seal assemblies 13760, which forces thefloating bushing 13766 against one of the stationary bushings 13764,13765 and forming a metal to metal seal 13769. At some point during eachstroke of the piston rod 13744, the pressure difference between theworkspace 13738 and the airlock 13736 will reverse. At about the sametime, the pressure difference reverses, the axial pressure across eachfloating rod seal assembly 13760 is zero. The floating bushing 13766 maymove radially within floating rod seal assembly 13760, when the pressuredifference is near zero and self-centered on the piston rod 13744. Adelay between a zero pressure difference across the piston rod seal unit13750 and zero axial pressure difference across the floating rod sealassembly 13760 may be due to flow resistances and volumes within/acrossthe piston rod seal unit 13750. As the piston rod 13744 continues itsstroke, the pressure difference across the floating rod seal assembly13760 reforms in the opposite direction and forces the floating bushing13766 against the other bushing 13764 or 13765, reforming a metal seal.Thus, the floating bushing 13766 may move radially during part of eachstroke to accommodate changes in the motion of the piston rod 13744 andform a seal against one of the bushings 13764, when significant pressuredifference occurs between the working space 13738 and the airlock 13736.

Referring now to FIG. 69D, in some embodiments, the piston rod seal unit13750 may include one or more floating clearance bushing 13870 that mayprovide a clearance seal with the piston rod 13744. The clearance sealis a long and narrow radial gap between the outside diameter of thepiston rod 13744 and the inside diameter of the floating clearancebushing 13870 that creates enough flow resistance that an insignificantamount of working gas leaks past the clearance seal. The oscillatingpressure of the workspace about the average pressure of the air lockassures that any gas that leaks from the workspace when the workspacepressure is relative to the air lock will leak back to the air lock whenthe workspace pressure is low.

In some embodiments, the inside diameter of the floating clearancebushing 13870 is 0.0005 to 0.001 inch larger than the outside diameterof the piston rod 13744 at room temperature and the floating clearancebushing 13870 is 0.71 inches long. The piston rod outer diameter, insome embodiments, may increase, for example, by a few ten-thousands ofan inch, due to thermal expansion during operation, when the rod may beapproximately 30 to 70° C. above room temperature. In some embodiments,to seal the work space 13738 from the air lock 13736 (FIG. 68A), thefloating clearance bushing 13870 (FIG. 69D) may also form an axial sealto prevent gas from flowing around the outside diameter of the floatingclearance bushing 13870. When the pressure is higher in the workingspace than the airlock, the floating clearance bushings 1370 may formaxial face seals with end plate 13756 and the washer 13858. When theworking space pressure is lower than the air lock pressure, the floatingclearance bushings will form an axial face seals with the threadedfitting 13876 and the washer 13858.

In various embodiments, the outer rings 13862 may be sized to allow thefloating clearance bushings 13870 to move radially and minimize axialmovement. In some embodiments, the outer ring 13862 may have an innerdiameter that is, e.g., 0.03 inch, larger than the outside diameter ofthe floating clearance bushing 13870 and the axial length of the outerring 13862 may be e.g. 0.0002 to 0.0005 inch, longer than the floatingclearance bushing. In some embodiments the length of the outer rings13862 may be matched to the length of the floating clearance bushing13870. The floating clearance bushings 13870, washer 13858, outer rings13862 may be assembled into the housing 13754 and this assembly may beaxially held in place between the end plate 13756 and the threadedfitting 13876. The threaded fitting 13876 threads into the housing 13754and contacts the outer rings 13862, which thereby set the axial spacingfor the floating clearance bushings 13870. The embodiment in FIG. 69Dmay include the particle trap 13780, scrapper ring 13778, port andfilter described for FIG. 69A. In other embodiments, the two floatingclearance bushings 13870 and washer 13858 in FIG. 69D may be replacedwith a single floating clearance bushing to reduce part count and cost.The single floating clearance bushing may in some examples be as long astwo of the floating clearance bushings 13870 in FIG. 69D.

Referring now to FIG. 69E, in some embodiments, the piston rod seal unit13750 includes one or more floating clearance bushings 13970 and a faceseal 13972 on one end of each floating clearance bushing 13970, wherethe floating clearance bushing 13970 provides a clearance seal with thepiston rod 13744. The face seals 13972 may take many forms including butnot limited to energized lip seals or o-rings. The face seals 13972 mayprovide improved axial sealing of the floating clearance bushing 13970to the threaded fitting 13876 and end plate 13756. The use of the faceseal may be beneficial for many reasons, including but not limited to,it may allow use of rougher surface finishes on the ends of the bushingsand mating surfaces. The embodiment in FIG. 69E may include the particletrap 13780, scrapper ring 13778, port and filter described for FIG. 69Aand the outer rings 13862, washer 13858, and housing 13754 described inFIG. 69D.

Referring now to FIG. 69F, in some embodiments, the piston rod seal unit13750 may include a hybrid seal that includes a floating clearancebushing 13971 and an energized lip seal 13752. In some embodiments, thehybrid seal may provide better sealing and a longer “life” orperformance time than either a clearance seal or a lip seal could bythemselves. The lip seal 13752 may provide a better seal than aclearance seal as the lip seal 13752 contacts the moving piston rod13744. The clearance seal between the floating clearance bushing 13971and the piston rod may reduce the pressure drop across the lip seal13752, which may extend the “life” or performance time of the lip seal.The floating seals 13971 may include face seals 13972 in someembodiments. The embodiment in FIG. 69F may include the particle trap13780, scrapper ring 13778, port and filter described for FIG. 69A andthe outer rings 13862, washer 13858, and housing 13754 described in FIG.69D.

The floating clearance bushings in FIGS. 69D-69F in various embodimentsmay be formed from a material with a low coefficient of friction, lowwear and high strength. Materials for floating clearance bushingsinclude, but are not limited to, one or more of the following: PTFE,Rulon, engineered plastics, graphite and graphite blends. In someembodiments, the material may be graphite impregnated with antimony, forexample, from the SGL Carbon Company of Germany produces graphiteimpregnated with antimony under the grade EK3205.

Linear Cross-Head Bearing

Referring now to FIG. 71A, which shows a cross-section of an embodimentof a Stirling engine with a drive 13920 to convert the linear motion ofthe piston 13930 into rotary motion of the crankshaft and generator.Drive 13920 is one example of a drive to convert linear motion of thepistons to rotary motion of the generator. Another example of a drive toconvert linear motion of the pistons to rotary motion of the generatoris in U.S. Pat. No. 6,253,550, which is incorporated by reference.

Still referring to FIG. 71A, the piston 13930 is fully guided by across-head in the form of linear bearing 13946. The linear cross-headbearing 13946 receives all the side loads in the lubricated drive of theengine and removes all side loads from seals on the piston and pistonrod. The lubricated linear cross-head bearing may absorb the side loadswith little friction and near zero wear. The linear cross-head bearing13946 has a sufficiently large length/diameter ratio, sufficient smallradial clearance and/or sufficient accurate alignment with the heaterhead to assure that the piston 13930 will not contact cylinder walls ofthe heater head 13712 nor the cooler 13915 during operation of theengine. In addition, the linear cross-head bearing 13946 may receiveside loads from the link rod 13719. The linear cross-head bearing 13946transfers these side loads to the engine block 13949 without negligiblewear or acceptable friction loads.

The linear cross-head bearing 13946 in FIGS. 71A-71C is configured toabsorb all the side loads from the link rod 13719 and fully guide thepiston 13930 and piston rod 13744 without creating side loads on thepiston seals 13934 or piston rod seal unit 13750. The linear cross-headbearing 13946 is preferably a hydrodynamic bearing that is fedpressurized oil. In one example, pressurized oil is feed into thebearing 13946 at annulus 13948 from passages (not shown) in the engineblock 13949. The ratio of length over the diameter of the linear bearing13946 is at least 2.0 In a preferred embodiment, length over diameterratio is 2.20. The diameter of the linear cross-head bearing 13946 has adiameter of 1.45 inches or more than 63% of the heater head 13712 innerdiameter. The selection of the linear bearing diameter balances axialstiffness which increases with diameter and frictional losses which alsoincrease with diameter. Other embodiments of the linear cross-headbearing may have different diameters. The diametrical gap of the linearcross-head bearing 13946 is generally equal to the diameter of thebearing divided by 1000. In a preferred embodiment the nominaldiametrical gap for the linear cross-head bearing is 0.0015 inches. Theaxial alignment of the linear cross-head bearings 13746 with the heaterhead 13712, cooler 13915 and piston rod seal unit 13750 is improved byboring the 4 guide holes for the 4 linear cross-head bearings 13946 in asingle setup. The 4 guide holes may be linebored to provide holes thatstraight and perpendicular to the engine block face 13949A.

Referring now to FIG. 14, the linear cross-head bearing removes the needfor guide ring 1616 on the piston 1600. The guide ring 1616 can be awear component that, in some instances, has a shorter life than thepressure seals 1606, 1614. The guide rings also occupy axial space onthe pistons that could be used for additional pressure seals 1606 1614.In some instances, the life of the pressure seals 1606, 1614 increaseswith the number of seals used on a given piston.

Referring now to FIG. 71B which presents a cross-sectional view of asingle piston/cylinder in a multi-cylinder engine and a portion of thedrive assembly. The linear cross-head bearing 13946 fully guides themotion of the piston rod 13744 and the piston assembly 13930. The linearcross-head bearing 13946 constrains the piston/piston rod to linearmovement with minimal rotation so that the piston 13930 cannot contactthe cylinder walls 13926. In one embodiment, the linear cross-sectionbearing 13946 has a minimum engaged length of 3.2 inches and adiametrical clearance of 0.0015 inches, so that the maximum radialmovement of the piston base is less than 0.005 inches which is half ofthe radial gap between the piston base 13932 and the cylinder wall13926.

In one embodiment, the linear cross-head bearing 13946 comprises ajournal or shaft 13947 that rides on a hydrodynamic layer of oil insidea hole or guide 13950 within the engine block 13949. The journal 13947may be connected at a first end via a rotatable joint to the link rod13719 which in turn is connected to the rocking beam 13916 (FIG. 71A).The journal 13947 may be rigidly connected on the second end to thepiston rod 13744. An oil pump (not shown) supplies pressurized oil tothe gap between the journal 13947 and the guide 13950 via oil annulus13948 in guide 13950. In a preferred embodiment the oil annulus islocated approximately equal distant from both ends of the guide 13950.The diametrical gap between the journal 13947 and the guild 13950 may bea value between 0.001 inch and 0.002 inch. In a preferred embodiment,the diametrical gap is 0.0015 inch. The guide 13950 is the section ofthe hole through the engine block 13949 that has a nearly constantdiameter that is slightly larger than the journal diameter. In apreferred embodiment the guide diameter is 0.001 to 0.002 inch largerthan the journal diameter. The linear cross-head bearing 13946 islocated below the piston, the piston rod seals and the bellows thatisolated the oil lubricated drive from the air lock and the workingspace.

Referring now to FIG. 71C, in some embodiments the second end of thejournal 13947 comprises an inner mount and inner support a rollingdiaphragm or bellows 13960 and a mounting point for the piston rod13744. In one example, the inner bead of the bellows 13960 is compressedto create a gas tight seal by washer 13966 that is held in place by ahollow screw 13964 that threads into the top of the journal 13947 whilethe piston rod 13744 pass through the hollow screw 13964. The outer beadof the bellows 13960 is trapped between collar 13962 and bellows support13968 that is mounted in the engine block 13949. In normal operation,the pressure in air lock 13766 is maintained above the crankcasepressure creating a positive pressure difference. The positive pressureforms the bellows as pictured in FIG. 71C. The positive pressuredifference presses the bellows 13960 against the outside diameter of thejournal 13947 and the inside diameter of the support 13968, so thatstress on the bellows is generally limited to a small unsupported Ushape. In use, the pressure difference between the airlock 13736 and thecrankcase may reverse so that the higher crank case will push thebellows 13960 up into the airlock space. The collar 13962 is sized andincludes a smooth lower corner to support the bellows 13960 out theoutside diameter in this inverted position which will reduce the stresson the bellows and the likelihood of bellows failure. In a preferredembodiment, the collar 13962 has an inner diameter that matches theinner diameter of the bellows support 13968 and a lower corner radius of0.10 inches.

Referring now to FIG. 72, one embodiment of the piston 13930 includes aclearance seal 13938 that provides pressure isolation between theworkspace above and below the piston without contacting the cylinderwalls. A no-contact seal would not wear and would have a much longer oreven infinite life. As the piston seals can the first element of aStirling engine to wear out or fail, a clearance seal could dramaticallyincrease the life and reduce the time between failures or maintenanceevents of the whole Stirling engine.

The outside diameter of the clearance seal 13938 nearly equal, but justsmaller than the inside diameter of the mating cylinder 13926 (in FIG.71B). In one embodiment, the clearance seal 13938 is fabricated from oneor more non-metal materials including but not limited to graphite, PTFE,Ultra High Molecular Weight Polyethylene (UHMWPE), ceramic, compositescontaining graphite, composites containing antimony, compositescontaining PTFE, composites containing UHMWPE. The cylinder innerdiameter may have a hard surface composed of material including but notlimited to tool steel, stainless steel, ceramic coatings. In oneembodiment, the outer diameter of the clearance seal 13938 is 0.0005 to0.001 inch smaller than the matting cylinder inner diameter. In anotherembodiment, the outer diameter of the clearance seal 13938 is machinedto be nominally equal to the cylinder inner diameter. The piston 13930is assembled into the cylinder and operated to wear down interferenceson the softer clearance seal 13938.

In some embodiments, the piston clearance seal 13938 may be graphite ora graphite impregnated with antimony and or may be made of any materialand/or any combination and/or composite of material. In someembodiments, the clearance seal 13938 may be made of antimonyimpregnated graphite which may be from the SGL Carbon Company ofGermany.

In one embodiment, the clearance seal 13938 includes a plurality ofradial grooves 13938A in the outer diameter. The grooves increase axialflow resistance in the gap between the clearance seal 13938 and thecylinder wall. One theory, among others, suggests that the repeatedgrooves 13938A perpendicular to axial flow disrupt that flow andincrease the flow resistance by creating repeated expansions andcontractions. In one embodiment the clearance seal 13938 is mounted inthe piston base 13932 and held in place by the piston dome 13934. Gas isprevented from leaking between the clearance seal and the piston base byone or more o-rings 13958.

The piston 13930 comprises a piston base 13942 that mounts to the pistonrod 13744, the piston dome 13934 and one or more seals 13938. The pistonmay also comprise one or more guide rings 13940 that may be required insome embodiments of the engine. In other embodiments of the piston13930, the guide rings 13940 are not included. The piston dome 13934 mayfurther comprise a inner dome 13936 that reduces the transfer of thermalenergy from the closed end of the piston dome 13934 to the piston base13942. The overall engine efficiency is improved by minimizing parasiticheat transfer from the hot end to the cold end of the engine. Parasiticheat transfer is the transfer of thermal energy from the hot end to thecold end that does not involve the working fluid. One example ofparasitic heat transfer is axial heat transfer from the hot closed endof the piston dome 13934 that is near the temperature of the heater tothe cold piston base 13932 that is near the temperature of the cooler.Therefore, reducing axial heat transfer through the piston will increasethe engine efficiency. In addition, the contact seals 13934 (FIG. 71B)last longer if they are kept cooler, so decreasing the axial heattransfer through the piston will increase the life of the contact seals.One theory, among others, states the inner dome 13936 reduces axial heattransfer as a radiation shield between the closed end of the piston dome13934 and the piston base 13932. Another theory states inner domereduces axial heat transfer by reducing convection in the dome createdin part by the oscillating movement of the piston. The mass of gas inthe piston dome 13934 and inner dome 13936 changes during engine startupand during some transient conditions. A port 13944 in the inner dome13936 and a small orifice 13942 in the piston base 13932 allow the gasto move into and out of the piston 13930. The orifice 13942 is smallenough to act as a low pass filter on the pressure in the piston. Theworking pressure below the piston base 13932 may change by more than onehundred psi every engine rotation. The orifice 13942 is small enough tokeep the pressure in the piston nearly constant over an engine rotationor cycle. In a preferred embodiment, the orifice 13942 has an 0.004 inchdiameter. The pressures in the inner dome and the outer dome do notoscillate, so the port 13944 between the two volumes can be ten timeslarger.

In some Stirling engines, reciprocating seals on piston are wear itemsthat wear away over time with normal use. The life of these piston sealsdetermines the time between failures or the time between maintenanceevents of the entire engine. A long time between maintenance events anda long life are important to the economic success of the Stirling. Thetime between maintenance or failures could be dramatically increased ifthe piston seals and piston rod seals were not wear items such asclearance seals. The decreased maintenance would increase the value andutility of the Stirling engine. An embodiment of the Stirling enginewithout non-wearing seals comprises an oil lubricated drive 13920 (FIG.71A) and the linear cross-head bearing 13946 (FIG. 71A-C), a clearanceseal 13938 (FIG. 72) on the piston and a piston rod seal unit 13750 withclearance seals 13862 (FIG. 69D, FIG. 69E). This embodiment with an oillubricated drive and non-contact reciprocating seals may have a timebetween maintenance events that is many times longer than a Stirlingengine with standard sliding seals.

In an embodiment of the Stiring engine where the piston seal and orpiston rod seal are contact seals, the linear cross-head bearingimproves the life of the seals and extends the time between failures orthe time between maintenance events for the engine. The piston andpiston rod are fully guided or constrained by the linear cross-headbearing so the possible radial movement of the piston and or piston rodis limited to an acceptably small value as described above. In addition,because the linear cross-head bearing does not wear, the radial movementof the piston and piston rod does not change over time. The reducedradial movement limits the loads applied to the piston seals and thepiston rod seals and the reduce loads increases the life of the theseseals.

Lubricating Fluid Pump and Lubricating Fluid Passageways

In some embodiments, the lubricating fluid is oil. The lubricating fluidis used to lubricate engine parts in the crankcase 2206, such ashydrodynamic pressure fed lubricated bearings. Lubricating the movingparts of the engine 2200 serves to further reduce friction betweenengine parts and further increase engine efficiency and engine life. Insome embodiments, lubricating fluid may be placed at the bottom of theengine, also known as an oil sump, and distributed throughout thecrankcase. The lubricating fluid may be distributed to the differentparts of the engine 2200 by way of a lubricating fluid pump, wherein thelubricating fluid pump may collect lubricating fluid from the sump via afiltered inlet. In the exemplary embodiment, the lubricating fluid isoil and thus, the lubricating fluid pump is herein referred to as an oilpump. However, the term “oil pump” is used only to describe theexemplary embodiment and other embodiments where oil is used as alubricating fluid, and the term shall not be construed to limit thelubricating fluid or the lubricating fluid pump.

Tube Heat Exchanger

External combustion engines, such as, for example, Stirling cycleengines, may use tube heater heads to achieve high power. FIG. 19 is across-sectional view of a cylinder and tube heater head of anillustrative Stirling cycle engine. A typical configuration of a tubeheater head 4200, as shown in FIG. 19, uses a cage of U-shaped heatertubes 4202 surrounding a combustion chamber 4204. A cylinder 4206contains a working fluid, such as, for example, helium. The workingfluid is displaced by the piston 4208 and driven through the heatertubes 4202. A burner 4210 combusts a combination of fuel and air toproduce hot combustion gases that are used to heat the working fluidthrough the heater tubes 4202 by conduction. The heater tubes 4202connect a regenerator 4212 with the cylinder 4204. The regenerator 2812may be a matrix of material having a large ratio of surface to areavolume which serves to absorb heat from the working fluid or to heat theworking fluid during the cycles of the engine. Heater tubes 2802 providea high surface area and a high heat transfer coefficient for the flow ofthe combustion gases past the heater tubes 4202. Various embodiments oftube heater heads are discussed below, and in U.S. Pat. No. 6,543,215and U.S. Pat. No. 7,308,787, which are, as previously mentioned,incorporated by reference in their entireties.

FIG. 19 is a side view in cross section of a tube heater head and acylinder. The heater head 4206 is substantially a cylinder having oneclosed end 4220 (otherwise referred to as the cylinder head) and an openend 4222. Closed end 4220 includes a plurality of U-shaped heater tubes4204 that are disposed in a burner 4200. Each U-shaped tube 4204 has anouter portion 4216 (otherwise referred to herein as an “outer heatertube”) and an inner portion 4218 (otherwise referred to herein as an“inner heater tube”). The heater tubes 4204 connect the cylinder 4202 toregenerator 4210. Cylinder 4202 is disposed inside heater head 4206 andis also typically supported by the heater head 4206. A piston 4224travels along the interior of cylinder 4202. As the piston 4224 travelstoward the closed end 4220 of the heater head 4206, working fluid withinthe cylinder 4202 is displaced and caused to flow through the heatertubes 4224 and regenerator 4210 as illustrated by arrows 4230 and 4232in FIG. 19. Referring to FIG. 19, as mentioned above, the closed end ofheater head 4220, including the heater tubes 4202, is disposed in aburner 4200 that includes a combustion chamber 4204. Hot combustiongases (otherwise referred to herein as “exhaust gases”) in combustionchamber 4204 are in direct thermal contact with heater tubes 4202 ofheater head 4220. Thermal energy is transferred by conduction from theexhaust gases to the heater tubes 4202 and from the heater tubes 4202 tothe working fluid of the engine, typically helium. Other gases, such asnitrogen, for example, or mixtures of gases, may be used, with apreferable working fluid having high thermal conductivity and lowviscosity. Non-combustible gases are used in various embodiments. Heatis transferred from the exhaust gases to the heater tubes 4202 as theexhaust gases flow around the surfaces of the heater tubes 4202. Arrows4230 show the general radial direction of flow of the exhaust gases. Theexhaust gases exiting from the burner 4210 tend to overheat the upperpart of the heater tubes 4202 (near the U-bend) because the flow of theexhaust gases is greater near the upper part of the heater tubes than atthe bottom of the heater tubes (i.e., near the bottom of the burner4200).

The overall efficiency of an external combustion engine is dependent inpart on the efficiency of heat transfer between the combustion gases andthe working fluid of the engine.

Returning to FIG. 19, in general, the inner heater tubes 4218 are warmerthan the outer heater tubes 4216 by several hundred degrees Celsius. Theburner power and thus the amount of heating provided to the workingfluid is therefore limited by the inner heater tube 4218 temperatures.The maximum amount of heat will be transferred to the working gas if theinner and outer heater tubes are nearly the same temperature. Generally,embodiments, as described herein, either increase the heat transfer tothe outer heater tubes or decrease the rate of heat transfer to theinner heater tubes.

An alternative embodiment of flow diverter fins is shown in FIG. 17.FIG. 17 is a top view of a section of a tube heater head includingsingle flow diverter fins in accordance with an embodiment. In thisembodiment, a single flow diverter fin 4002 is connected to each outerheater tube 4004. In some embodiments, the flow diverter fins 4002 areattached to an outer heater tube 4004 using a nickel braze along thefull length of the heater tube. Alternatively, the flow diverter finsmay be brazed with other high temperature materials, welded or joinedusing other techniques known in the art that provide a mechanical andthermal bond between the flow diverter fin and the heater tube. Flowdiverter fins 4002 are used to direct the exhaust gas flow around theheater tubes 4004, including the downstream side of the heater tubes4004. In order to increase the heat transfer from the exhaust gas to theheater tubes 4004, flow diverter fins 4002 are thermally connected tothe heater tube 4004. Therefore, in addition to directing the flow ofexhaust gas, flow diverter fins 4002 increase the surface area for thetransfer of heat by conduction to the heater tubes 4004, and thence tothe working fluid.

FIG. 18 is a top view in cross-section of a section of a tube heaterhead including the single flow diverter fins as shown in FIG. 17 inaccordance with an embodiment. As shown in FIG. 18, a flow diverter fin4110 is placed on the upstream side of a heater tube 4106. The diverterfin 4110 is shaped so as to maintain a constant distance from thedownstream side of the heater tube 4106 and therefore improve thetransfer of heat to the heater tube 4106. In an alternative embodiment,the flow diverter fins could be placed on the inner heater tubes 4108.

Engine performance, in terms of both power and efficiency, is highest atthe highest possible temperature of the working gas in the expansionvolume of the engine. The maximum working gas temperature, however, istypically limited by the properties of the heater head. For an externalcombustion engine with a tube heater head, the maximum temperature islimited by the metallurgical properties of the heater tubes. If theheater tubes become too hot, they may soften and fail resulting inengine shut down. Alternatively, at too high of a temperature the tubeswill be severely oxidized and fail. It is, therefore, important toengine performance to control the temperature of the heater tubes. Atemperature sensing device, such as a thermocouple, may be used tomeasure the temperature of the heater tubes. The temperature sensormounting scheme may thermally bond the sensor to the heater tube andisolate the sensor from the much hotter combustion gases. The mountingscheme should be sufficiently robust to withstand the hot oxidizingenvironment of the combustion-gas and impinging flame that occur nearthe heater tubes for the life of the heater head. One set of mountingsolutions include brazing or welding thermocouples directly to theheater tubes. The thermocouples would be mounted on the part of theheater tubes exposed to the hottest combustion gas. Other possiblemounting schemes permit the replacement of the temperature sensor. Inone embodiment, the temperature sensor is in a thermowell thermallybonded to the heater tube. In another embodiment, the mounting scheme isa mount, such as a sleeve, that mechanically holds the temperaturesensor against the heater tube.

FIG. 19 is a side view in cross section of a cylinder 4204 and a burner4210. A temperature sensor 4202 is used to monitor the temperature ofthe heater tubes and provide feedback to a fuel controller (not shown)of the engine in order to maintain the heater tubes at the desiredtemperature. In some embodiments, the heater tubes are fabricated usingInconel 625 and the desired temperature is 930.degree. C. The desiredtemperature will be different for other heater tube materials. Thetemperature sensor 4202 should be placed at the hottest, and thereforethe limiting, part of the heater tubes. Generally, the hottest part ofthe heater tubes will be the upstream side of an inner heater tube 4206near the top of the heater tube. FIG. 19 shows the placement of thetemperature sensor 4202 on the upstream side of an inner heater tube4206. In some embodiments, as shown in FIG. 19, the temperature sensor4202 is clamped to the heater tube with a strip of metal 4212 that iswelded to the heater tube in order to provide good thermal contactbetween the temperature sensor 4202 and the heater tube 4206. In oneembodiment, both the heater tubes 4206 and the metal strip 4212 may beInconel 625 or other heat resistant alloys such as Inconel 600,Stainless Steels 310 and 316 and Hastelloy X. The temperature sensor4202 should be in good thermal contact with the heater tube, otherwiseit may read too high a temperature and the engine will not produce asmuch power as possible. In an alternative embodiment, the temperaturesensor sheath may be welded directly to the heater tube.

In another embodiment, as shown in FIG. 20A-B, a temperature sensormount 4320 is created with a formed strip or sheath of a refractory orhigh temperature resistant metal such as Inconel that is bonded to theexterior of the heater tube 4310. The sensor mount sheath 4320 is formedor shaped into a channel that when attached to the heater tube creates avoid that accommodates a device. In a specific embodiment, the channelis V-shaped to accommodate the insertion of a thermal sensor such as athermocouple device. The shaped channel is then bonded to the exteriorof a heater tube 4310 as shown in FIG. 20A.

FIG. 20A shows a side view of the sensor mount sheath 4320 on the heatertube 4310, while FIG. 20B is a view along the axis of the sensor mountsheath 4320. The metal should be thin enough to form, yet thick enoughto survive for the rated life of the heater head. In some embodiments,the metal is approximately between 0.005″ and 0.020″ thick. The metalmay be bent such that the bend is along the length of the strip. This“V-channel” sheath 4320 is then affixed to the exterior of the heatertube by high temperature brazing. Prior to brazing, the sheath may betack welded in several places to insure that the sheath does not moveduring the brazing process, as shown in FIG. 20A. Preferably, the brazecompound used during brazing is typically a high nickel alloy; however,any compound which will withstand the brazing temperature will work.Alternatively the sheath may be bonded to the heater tube by electronbeam or laser welding.

Now referring to FIG. 20B, a cavity 4330 is formed by affixing thesheath to the heater tube. This cavity 4330 is formed such that it mayaccept a device such as a thermocouple. When formed and brazed, thecavity may advantageously be sized to fit the thermocouple. Preferably,the fit is such that the thermocouple is pressed against the exterior ofthe heater tube. Preferably, the sheath is thermally connected to theheater tube. If the sheath is not thermally connected to the heatertube, the sheath may not be “cooled” by the working gas. The lack ofcooling may cause the sheath to operate at or near the combustion gastemperatures, which are typically high enough to eventually burn throughany metal. Brazing the sensor mount to the heater tube leads to a goodthermal contact. Alternatively, the sensor mount sheath 4320 could becontinuously welded along both sides to provide sufficient thermalconnection.

In another embodiment, as shown in FIGS. 21A-B, a second strip of metalcan be formed to create a shield 4450 over the sensor mount 4420. Theshield 4420 may be used to improve the thermal connection between thetemperature sensor, in cavity 4430, and the heater tube 4410. The shieldinsulates the sensor mount sheath 4420 from the convective heating ofthe hot combustion gases and thus improves the thermal connection to theheater tube. Furthermore, there is preferably an insulating space 4440to help further insulate the temperature sensor from the hot combustiongases as shown in FIG. 21B.

In another specific embodiment, as shown in FIGS. 22A and 22B, thetemperature sensor mount 4520 can be a small diameter tube or sleeve4540 joined to the leading edge of the heater tube 4510. FIG. 22A showsa side view of the mount on the heater tube 4510, while FIG. 22B is aview along the axis of the tube 4540 or sleeve. The sensor tube 4540 ispreferably brazed to the heater tube with a substantial braze fillet4530. The large braze fillet 4530 will maximize the thermal bond betweenthe heater tube and the sensor mount. In another embodiment, the tube orsleeve 4540 may have a shield. As described supra, an outer shield covermay help insulate the temperature sensor mount 4520 from convective heattransfer and improve the thermal connection to the heater tube.

In some embodiments of the heater head, inserts may be placed on theinside of the heater tubes to increase heat transfer between the workinggas and the tube walls as disclosed in U.S. patent application Ser. No.13/447,990, filed Apr. 16, 2012 and entitled Stirling Cycle Machine(Attorney Docket No. 184), which is hereby incorporated herein byreference in its entirety.

Regenerator

A regenerator is used in a Stirling cycle machine, as discussed aboveand as described in U.S. Pat. No. 6,591,609, and U.S. Pat. No.6,862,883, to add and remove heat from the working fluid duringdifferent phases of the Stirling cycle. The regenerator used in aStirling cycle machine must be capable of high heat transfer rates whichtypically suggests a high heat transfer area and low flow resistance tothe working fluid. Low flow resistance also contributes to the overallefficiency of the engine by reducing the energy required to pump theworking fluid. Additionally, a regenerator must be fabricated in such amanner as to resist spalling or fragmentation because fragments may beentrained in the working fluid and transported to the compression orexpansion cylinders and result in damage to the piston seals.

One regenerator design uses several hundred stacked metal screens. Whileexhibiting a high heat transfer surface, low flow resistance and lowspalling, metal screens may suffer the disadvantage that their cuttingand handling may generate small metal fragments that must be removedbefore assembling the regenerator. Additionally, stainless steel wovenwire mesh contributes appreciably to the cost of the Stirling cycleengine.

A three dimensional random fiber network, such as stainless steel woolor ceramic fiber, for example, may be used as the regenerator, as nowdescribed with reference to FIG. 23A. Stainless steel wool regenerator6300 advantageously provides a large surface area to volume ratio,thereby providing favorable heat transfer rates at low fluid flowfriction in a compact form. Additionally, cumbersome manufacturing stepsof cutting, cleaning and assembling large numbers of screens areadvantageously eliminated. The low mechanical strength of steel wool andthe tendency of steel wool to spall may both be overcome as nowdescribed. In some embodiments, the individual steel wires 6302 and 6304are “cross-linked” into a unitary 3D wire matrix.

The starting material for the regenerator may be fibrilose and of randomfiber form such as either steel or nickel wool. The composition of thefiber may be a glass or a ceramic or a metal such as steel, copper, orother high temperature materials. The diameter of the fiber ispreferably in the range from 10 micrometers to 1 millimeter depending onthe size of the regenerator and the properties of the metal. Thestarting material is placed into a form corresponding to the final shapeof the regenerator which is depicted in cross-section in FIG. 23B. Innercanister cylindrical wall 6320, outer canister cylindrical wall 6322,and regenerator network 6300 are shown. The density of the regeneratoris controlled by the amount of starting material placed in the form. Theform may be porous to allow fluids to pass through the form.

In some embodiments, unsintered steel wool is employed as regeneratornetwork 6300. Regenerator network 6300 is then retained within theregenerator canister by regenerator retaining screens 6324 or otherfilter, thereby comprising a “basket” which may advantageously capturesteel wool fragments.

In yet other embodiments, knit or woven wire is employed in fabricationof a regenerator as now described with reference to FIG. 24A. Inaccordance with these embodiments, knit or woven wire tube 6401 isflattened by rollers 6402 into tape 6404, in which form it is woundabout mandrel 6406 into annular layers 6408. Stainless steel isadvantageously used for knit wire tube 6401 because of its ability towithstand elevated temperature operation, and the diameter of the wireused is typically in the range of 1-2 mils, however other materials andgauges may be used in various embodiments. Alternatively, a plurality,typically 5-10, of the stainless steel wires may be loosely wound into amulti-filament thread prior to knitting into a wire tube. This processadvantageously strengthens the resulting tube 6401. When mandrel 6406 isremoved, annular assembly 6410 may be used as a regenerator in a thermalcycle engine.

Still another embodiment is now described with reference to FIGS. 24Bthrough 24E. Knit or woven wire tube 6401, shown in its rightcylindrical form in FIG. 24B, is shown scored and partially compressedin FIG. 24C. Alternatively, the scoring may be at an angle 6414 withrespect to the central axis 6412 of the tube, as shown in FIG. 24D. Tube6401 is then axially compressed along central axis 6412 to form thebellows form 6416 shown in FIG. 24E that is then disposed as aregenerator within the regenerator volume 408 (shown in FIG. 4) of aStirling cycle engine.

It is to be understood that the various regenerator embodiments andmethods for their manufacture described herein may be adapted tofunction in a multiple cylinder configuration.

Coolant Penetrating Cold-End Pressure Vessel

Referring now to FIGS. 25A-C, various cross-sections of an engine, suchas a Stirling cycle engine, are shown in accordance with someembodiments. Engine 6500 is hermetically sealed. A crankcase 6502 servesas the cold-end pressure vessel and contains a charge gas in an interiorvolume 6504. Crankcase 6502 can be made arbitrarily strong withoutsacrificing thermal performance by using sufficiently thick steel orother structural material. A heater head 6506 serves as the hot-endpressure vessel and is preferably fabricated from a high temperaturesuper-alloy such as Inconel 625, GMR-235, etc. Heater head 6506 is usedto transfer thermal energy by conduction from an external thermal source(not shown) to the working fluid. Thermal energy may be provided fromvarious heat sources such as solar radiation or combustion gases. Forexample, a burner, as previously discussed, may be used to produce hotcombustion gases (shown as 6507 in FIG. 25B) that are used to heat theworking fluid. An expansion area of cylinder (or warm section) 6522 isdisposed inside the heater head 6506 and defines part of a working gasvolume as discussed above with respect to FIG. 1. A piston 6528 is usedto displace the working fluid contained in the expansion area ofcylinder 6522.

In accordance with an embodiment, crankcase 6502 is welded directly toheater head 6506 at joints 6508 to create a pressure vessel that can bedesigned to hold any pressure without being limited, as are otherdesigns, by the requirements of heat transfer in the cooler. In analternative embodiment, the crankcase 6502 and heater head 6506 areeither brazed or bolted together. The heater head 6506 has a flange orstep 6510 that axially constrains the heater head and transfers theaxial pressure force from the heater head 6506 to the crankcase 6502,thereby relieving the pressure force from the welded or brazed joints6508. Joints 6508 serve to seal the crankcase 6502 (or cold-end pressurevessel) and bear the bending and planar stresses. In an alternativeembodiment, the joints 6508 are mechanical joints with an elastomerseal. In yet another embodiment, step 6510 is replaced with an internalweld in addition to the exterior weld at joints 6508.

Crankcase 6502 is assembled in two pieces, an upper crankcase 6512 and alower crankcase 6516. The heater head 6506 is first joined to the uppercrankcase 6512. Second, a cooler 6520 is installed with a coolant tubing(shown as 6514 in FIG. 25B) passing through holes in the upper crankcase6512. Third, the double acting pistons 6528 and drive components(designated generally as numeral 6540 in FIGS. 25A and 25C, not shown inFIG. 25B) are installed. In one embodiment, lower crankcase 6516 isassembled in three pieces, an upper section 6513, a middle section 6515,and a lower section 6517, as shown in FIGS. 25A and 25C. Middle section6515 is may be connected to upper and lower sections 6513 and 6517 atjoints 6519 and 6521, respectively, by any mechanical means known in theart, or by welding.

The lower crankcase 6516 is then joined to the upper crankcase 6512 atjoints 6518. Preferably, the upper crankcase 6512 and the lowercrankcase 6516 are joined by welding. Alternatively, a bolted flange maybe employed (as shown in FIGS. 25B and 25C).

In some embodiments a motor/generator (shown as 6501 in FIG. 25C), suchas a PM generator, may be installed into motor/generator housing (shownas 6503 in FIG. 25C), which is attached to the lower crankcase 6516, asshown in FIG. 25C. Motor/generator housing 6503 may be attached to lowercrankcase 6516 by any mechanical means known in the art, or may bewelded to lower crankcase 6516. Motor/generator housing 6503 mayassembled in two pieces, a front section 6505, which is attached tolower crankcase 6516, and a rear section 6509, which may be welded orbolted to front section 6505. In one embodiment a seal 6511 may bepositioned between the rear section 6509 and the front section 6505 ofthe motor/generator housing 6503. In some embodiments rear section 6509is removable attached to front section 6505, which serves, among otherfunctions, to allow for easy removal and installation of motor/generator6501 during engine 6500 assembly.

In order to allow direct coupling of the heater head 6506 to the uppercrankcase 6512, the cooling function of the thermal cycle is performedby a cooler 6520 that is disposed within the crankcase 6502, therebyadvantageously reducing the pressure containment requirements placedupon the cooler. By placing the cooler 6520 within crankcase 6502, thepressure across the cooler is limited to the pressure difference betweenthe working gas in the working gas volume, and the charge gas in theinterior volume 6504 of the crankcase. The difference in pressure iscreated by the compression and expansion of the working gas, and istypically limited to a percentage of the operating pressure. In oneembodiment, the pressure difference is limited to less than 30% of theoperating pressure.

Coolant tubing 6514 advantageously has a small diameter relative to thediameter of the cooler 6520. The small diameter of the coolant passages,such as provided by coolant tubing 6514, is key to achieving high heattransfer and supporting large pressure differences. The required wallthickness to withstand or support a given pressure is proportional tothe tube or vessel diameter. The low stress on the tube walls allowsvarious materials to be used for coolant tubing 6514 including, but notlimited to, thin-walled stainless steel tubing or thicker-walled coppertubing.

An additional advantage of locating the cooler 6520 entirely within thecrankcase 6502 (or cold-end pressure vessel) volume is that any leaks ofthe working gas through the cooler 6520 will only result in a reductionof engine performance. In contrast, if the cooler were to interface withthe external ambient environment, a leak of the working gas through thecooler would render the engine useless due to loss of the working gasunless the mean pressure of working gas is maintained by an externalsource. The reduced requirement for a leak-tight cooler allows for theuse of less expensive fabrication techniques including, but not limitedto, powder metal and die casting.

Cooler 6520 is used to transfer thermal energy by conduction from theworking gas and thereby cool the working gas. A coolant, either water oranother fluid, is carried through the crankcase 6502 and the cooler 6520by coolant tubing 6514. The feedthrough of the coolant tubing 6514through upper crankcase 6512 may be sealed by a soldered or brazed jointfor copper tubes, welding, in the case of stainless steel and steeltubing, or as otherwise known in the art.

The charge gas in the interior volume 6504 may also require cooling dueto heating resulting from heat dissipated in the motor/generatorwindings, mechanical friction in the drive, the non-reversiblecompression/expansion of the charge gas, and the blow-by of hot gasesfrom the working gas volume. Cooling the charge gas in the crankcase6502 increases the power and efficiency of the engine as well as thelongevity of bearings used in the engine.

In one embodiment, an additional length of coolant tubing (shown as 6530in FIG. 25B) is disposed inside the crankcase 6502 to absorb heat fromthe charge gas in the interior volume 6504. The additional length ofcoolant tubing 6530 may include a set of extended heat transfer surfaces(shown as 6548 in FIG. 25B), such as fins, to provide additional heattransfer. As shown in FIG. 25B, the additional length of coolant tubing6530 may be attached to the coolant tubing 6514 between the crankcase6502 and the cooler 6520. In an alternative embodiment, the length ofcoolant tubing 6530 may be a separate tube with its own feedthrough ofthe crankcase 6502 that is connected to the cooling loop by hosesoutside of the crankcase 6502.

In another embodiment the extended coolant tubing 6530 may be replacedwith extended surfaces on the exterior surface of the cooler 6520 or thedrive housing (shown as 6572 in FIGS. 25A and 25C). Alternatively, a fan(shown as 6534 in FIG. 25B) may be attached to the engine crankshaft(shown as 6542 in FIG. 25C) to circulate the charge gas in interiorvolume 6504. The fan 6534 may be used separately or in conjunction withthe additional coolant tubing 6530 or the extended surfaces on thecooler 6520 or drive housing 6572 to directly cool the charge gas in theinterior volume 6504.

Preferably, coolant tubing 6514 is a continuous tube throughout theinterior volume 6504 of the crankcase and the cooler 6520.Alternatively, two pieces of tubing could be used between the crankcaseand the feedthrough ports of the cooler. One tube carries coolant fromoutside the crankcase 6502 to the cooler 6520. A second tube returns thecoolant from the cooler 6520 to the exterior of the crankcase 6502. Inanother embodiment, multiple pieces of tubing may be used between thecrankcase 6502 and the cooler in order to add tubing with extended heattransfer surfaces inside the crankcase volume 6504 or to facilitatefabrication. The tubing joints and joints between the tubing and thecooler may be brazed, soldered, welded or mechanical joints.

Various methods may be used to join coolant tubing 6514 to cooler 6520.Any known method for joining the coolant tubing 6514 to the cooler 6520may be used in various embodiments. In one embodiment, the coolanttubing 6514 may be attached to the wall of the cooler 6520 by brazing,soldering or gluing. Cooler 6520 is in the form of a cylinder placedaround the cylinder 6522 and the annular flow path of the working gasoutside of the cylinder 6522. Accordingly, the coolant tubing 6514 maybe wrapped around the interior of the cooler cylinder wall and attachedas mentioned above. Various embodiments of cooler configurations may befound in U.S. patent application Ser. No. 13/447,990, filed Apr. 16,2012 and entitled Stirling Cycle Machine (Attorney Docket No. 184),which is hereby incorporated herein by reference in its entirety.

Returning to FIG. 25B, one method for joining coolant tubing 6514 tocooler 6520 is to overcast the cooler around the coolant tubing. Thismethod is described, with reference to FIGS. 26A and 26B, and may beapplied to a pressurized close-cycle machine as well as in otherapplications where it is advantageous to locate a cooler inside thecrankcase.

Referring to FIG. 26A, a heat exchanger, for example, a cooler 6520(shown in FIGS. 25A and 25B) may be fabricated by forming ahigh-temperature metal tubing 6602 into a desired shape. In oneembodiment, the metal tubing 6602 is formed into a coil using copper. Alower temperature (relative to the melting temperature of the tubing)casting process is then used to overcast the tubing 6602 with a highthermal conductivity material to form a gas interface 6604 (and 6532 inFIG. 25B), seals 6606 (and 6524 in FIG. 25B) to the rest of the engineand a structure to mechanically connect the drive housing 6572 (shown inFIG. 25C) to the heater head 6506 (shown in FIG. 25B. In one embodiment,the high thermal conductivity material used to overcast the tubing isaluminum. Overcasting the tubing 6602 with a high thermal conductivitymetal assures a good thermal connection between the tubing and the heattransfer surfaces in contact with the working gas. A seal is createdaround the tubing 6602 where the tubing exits the open mold at 6610.This method of fabricating a heat exchanger advantageously providescooling passages in cast metal parts inexpensively.

FIG. 26B is a perspective view of a cooling assembly cast over thecooling coil of FIG. 26A. The casting process can include any of thefollowing: die casting, investment casting, or sand casting. The tubingmaterial is chosen from materials that will not melt or collapse duringthe casting process. Tubing materials include, but are not limited to,copper, stainless steel, nickel, and super-alloys such as Inconel. Thecasting material is chosen among those that melt at a relatively lowtemperature compared to the tubing. Typical casting materials includealuminum and its various alloys, and zinc and its various alloys.

The heat exchanger may also include extended heat transfer surfaces toincrease the interfacial area 6604 (and 6532 shown in FIG. 25B) betweenthe hot working gas and the heat exchanger so as to improve heattransfer between the working gas and the coolant. Extended heat transfersurfaces may be created on the working gas side of the heat exchanger6520 by machining extended surfaces on the inside surface (or gasinterface) 6604. Referring to FIG. 25B, a cooler liner 6526 (shown inFIG. 25B) may be pressed into the heat exchanger to form a gas barrieron the inner diameter of the heat exchanger. The cooler liner 6526directs the flow of the working gas past the inner surface of thecooler.

The extended heat transfer surfaces can be created by any of the methodsknown in the art. In accordance some embodiments, longitudinal grooves6704 are broached into the surface, as shown in detail in FIG. 27A.Alternatively, lateral grooves 6708 (also shown in enlarged section viewFIG. 27D) may be machined in addition to the longitudinal grooves 6704(also shown in enlarged section view FIG. 27B) thereby creating alignedpins 6710 as shown in FIG. 27C. In some embodiments, grooves are cut ata helical angle to increase the heat exchange area.

In an alternative embodiment, the extended heat transfer surfaces on thegas interface 6604 (as shown in 26B) of the cooler are formed from metalfoam, expanded metal or other materials with high specific surface area.For example, a cylinder of metal foam may be soldered to the insidesurface of the cooler 6604. As discussed above, a cooler liner 6526(shown in FIG. 25B) may be pressed in to form a gas barrier on the innerdiameter of the metal foam. Other methods of forming and attaching heattransfer surfaces to the body of the cooler are described in U.S. Pat.No. 6,694,731, issued Feb. 24, 2004, entitled Stirling Engine ThermalSystem Improvements, which is herein incorporated by reference in itsentirety.

Additional coolant penetrating cold-end pressure vessel embodiments aredescribed in U.S. Pat. No. 7,325,399. It is to be understood that thevarious coolant penetrating cold-end pressure vessel embodimentsreferred to herein may be adapted to function in a multiple cylinderengine configuration.

Intake Manifold

Referring now to FIGS. 28A-28C, an intake manifold 6899, is shown forapplication to a Stirling cycle engine or other combustion applicationin accordance with some embodiments. Various embodiments of intakemanifold 6899 are further disclosed in U.S. Pat. No. 6,381,958. Inaccordance with some embodiments, fuel is pre-mixed with air that may beheated above the fuel's auto-ignition temperature and a flame isprevented from forming until the fuel and air are well-mixed. FIG. 28Ashows one embodiment including an intake manifold 6899 and a combustionchamber 6810. The intake manifold 6899 has an axisymmetrical conduit6801 with an inlet 6803 for receiving air 6800. Air 6800 is pre-heatedto a temperature, typically above 900 K, which may be above theauto-ignition temperature of the fuel. Conduit 6801 conveys air 6800flowing inward radially with respect to combustion axis 6820 to aswirler 6802 disposed within the conduit 6801.

FIG. 28B shows a cross sectional view of the conduit 6801 includingswirler 6802 in accordance with some embodiments. In the embodiment ofFIG. 28B, swirler 6802 has several spiral-shaped vanes 6902 fordirecting the flow of air 6800 radially inward and imparting arotational component on the air. The diameter of the swirler section ofthe conduit decreases from the inlet 6904 to the outlet 6906 of swirler6802 as defined by the length of the swirler section conduit 6801. Thedecrease in diameter of swirler vanes 6902 increases the flow rate ofair 6800 in substantially inverse proportion to the diameter. The flowrate is increased so that it is above the flame speed of the fuel. Atoutlet 6906 of swirler 6802, fuel 6806, which in a one embodiment ispropane, is injected into the inwardly flowing air.

In some embodiments, fuel 6806 is injected by fuel injector 6804 througha series of nozzles 6900 as shown in FIG. 28C. More particularly, FIG.28C shows a cross sectional view of conduit 6801 and includes the fueljet nozzles 6900. Each of the nozzles 6900 is positioned at the exit ofthe swirler vanes 6902 and is centralized between two adjacent vanes.Nozzles 6900 are positioned in this way for increasing the efficiency ofmixing the air and fuel. Nozzles 6900 simultaneously inject the fuel6806 across the air flow 6800. Since the air flow is faster than theflame speed, a flame will not form at that point even though thetemperature of the air and fuel mixture is above the fuel'sauto-ignition temperature. In some embodiments, where propane is used,the preheat temperature, as governed by the temperature of the heaterhead, is approximately 900 K.

Referring again to FIG. 28A, the air and fuel, now mixed, referred tohereafter as “air-fuel mixture” 6809, is transitioned in directionthrough a throat 6808 which has a contoured fairing 6822 and is attachedto the outlet 6807 of the conduit 6801. Fuel 6806 is supplied via fuelregulator 6824.

Throat 6808 has an inner radius 6814 and an outer dimension 6816. Thetransition of the air-fuel mixture is from a direction which issubstantially transverse and radially inward with respect to combustionaxis 6820 to a direction which is substantially parallel to thecombustion axis. The contour of the fairing 6822 of throat 6808 has theshape of an inverted bell such that the cross sectional area of throat6808 with respect to the combustion axis remains constant from the inlet6811 of the throat to outlet 6812 of the throat. The contour is smoothwithout steps and maintains the flow speed from the outlet of theswirler to the outlet of the throat 6808 to avoid separation and theresulting recirculation along any of the surfaces. The constant crosssectional area allows the air and fuel to continue to mix withoutdecreasing the flow speed and causing a pressure drop. A smooth andconstant cross section produces an efficient swirler, where swirlerefficiency refers to the fraction of static pressure drop across theswirler that is converted to swirling flow dynamic pressure. Swirlefficiencies of better than 80% may typically be achieved in practice.Thus, the parasitic power drain of the combustion air fan may beminimized.

Outlet 6812 of the throat flares outward allowing the air-fuel mixture6809 to disperse into the chamber 6810 slowing the air-fuel mixture 6809thereby localizing and containing the flame and causing a toroidal flameto form. The rotational momentum generated by the swirler 6802 producesa flame stabilizing ring vortex as well known in the art.

Gaseous Fuel Burner

Definitions: As used in this section of the detailed description, thefollowing terms shall have the meanings indicated, unless the contextotherwise requires: Fuel-Air Equivalence ratio (.phi.)=Actual Fuel-AirMass Ratio/Stoichiometric Fuel-Air Mass Ratio. The stoichiometricfuel-air mass ratio is defined as the mass ratio needed to balance thefuel+air chemical equation. The stoichiometric fuel-air mass ratio iswell known for common fuels such as propane (0.0638 g fuel/g air) andcalculable for gases such as biogas.

FIG. 29 shows one embodiment of the engine 7212 embodiment having agaseous fuel burner 7201. Various embodiments of the gaseous fuel burner7201 are also disclosed in U.S. patent application Ser. No. 11/122,447,filed May 5, 2005, published Nov. 10, 2005, which is herein incorporatedby reference in its entirety. This embodiment may be used in the contextof a Stirling cycle engine, however, other embodiments of the machineare not limited to such applications. Those skilled in the art willappreciate that the present machine may have application in othersystems, such as, with other types of external combustion engines.

The use of an ejector in a gaseous fuel burner advantageously can solvesome of the challenges faced by the traditional gaseous fuel burners.First, using an ejector can eliminate the need for additional equipment,controls, and space, such as, a gaseous fuel pump, fuel controlcircuitry, and the associated components. Furthermore, using an ejectorsuch as a venturi simplifies the fuel control system by eliminating theneed for a separate fuel control scheme. Based on the corresponding riseof the vacuum with the airflow, and subsequently, an increased fuelflow, the burner power can be regulated by regulating the airflow.Accordingly, removing separate fuel control simplifies the developmentand implementation of automatic burner control in a gaseous fuel burnerwith an ejector.

Secondly, the corresponding rise of the vacuum with airflow also resultsin an approximately steady fuel-air ratio despite changes in temperatureand airflow rates. The resulting steady fuel-air ratio simplifies thefuel control and operation of the burner, by eliminating the need forcomplex exhaust sensor/feedback fuel control mechanisms.

Referring to FIG. 29, a gaseous fuel burner 7201 comprises an ejector7240, a heat exchanger 7220, a combustion chamber 7250, and a blower7200 (shown as 7300 in FIG. 30A). The term ejector as used here includeseductors, siphons, or any device that can use the kinetic energy of onefluid to cause the flow of another fluid. Ejectors are a reliable way ofproducing vacuum-based fuel flow systems with low initial cost, lack ofmoving parts, and simplicity of operation.

Referring again to FIG. 29, in a some embodiments, the ejector 7240 is aventuri. The venturi 7240 is positioned downstream of the outlet of theair preheater or heat exchanger 7220, in a venturi plenum 7241 andproximal to the combustion chamber 7250. A blower 7200 forces airthrough the venturi 7240. The flow of air through the venturi draws in aproportional amount of fuel through the fuel inlet ports 7279. The fuelinlet ports 7279 are placed at the venturi throat 7244 where the throathas the lowest pressure. The ports 7279 are sized to produce plumes offuel across the airflow that promote good mixing within the venturi7240. This fuel-air mixture exits the venturi 7240 and forms aswirl-stabilized flame in the combustion chamber 7250. The venturi 7240draws in an amount of fuel that is substantially linearly proportionalto the airflow regardless of airflow rates and temperature of the airentering the venturi 7240.

In a some embodiments as shown in FIGS. 30A and 30B, placing the venturi7340 between the air preheater 7320 and the combustion chamber 7350promotes a substantially steady air-fuel ratio over a wide range ofairflows and venturi temperatures. FIG. 30A is a schematic drawing ofthe burner including the components of the burner such as a blower 7300,a preheater 7320, a venturi 7340, and fuel supply 7372. The drawing alsoincludes a load heat exchanger or heater head 7390 (also shown in FIGS.31-33 as 7290). The load heat exchanger 7390 is the heat exchanger ofthe engine or process that absorbs the thermal power of the hot gasesleaving the combustion chamber 7350 in the burner at some elevatedtemperature. The partially cooled burned gases then enter the exhaustside of the air preheater, where they are further cooled by incomingcombustion air. FIG. 30B shows the pressure map of the same componentsarranged linearly. The air pressure supplied by the blower, the fuelsupply pressure, and the ambient pressure are all indicated. The massflow rate (m′) of fuel into the burner is controlled by the differencebetween the fuel supply pressure at 7372 and the pressure in the venturithroat 7344 (shown in FIG. 29 as 7244) and the fuel temperature at thedominant restriction:

-   -   m′.sub.FUEL.varies.(P.sub.FUEL-P.sub.THROAT).sup.0.5/T.sub.FUEL.sup.0.5

The pressure in the throat (P.sub.THROAT) is set by the pressure dropthrough the exhaust side of the preheater 7320 plus the pressure dropthrough the heater head tubes 7390 minus the suction generated by theventuri throat 7344. The pressure drops 7320, 7390 and the throatsuction pressure 7344 are all proportional to the airflow rate and theventuri temperature.

-   -   P.sub.THROAT.varies.m′.sub.AIR.sup.2*T.sub.VENTURI

Combining these equations shows that the fuel flow will varyapproximately linearly with the airflow:

m′.sub.FUEL.varies.[P.sub.FUEL-(m′.sub.AIR.sup.2*T.sub.VENTURI)].sup.0.5/T-.sub.FUEL.sup.0.5

Regulating the fuel pressure to near ambient pressure, the fuel flow isapproximately linear with airflow.

m′.sub.FUEL.varies.m′.sub.AIR*(T.sub.VENTURI/T.sub.FUEL).sup.0.5 Thus,locating the dominant fuel restriction 7378 (shown as 7278 in FIG. 29)within the venturi plenum (shown as 7241 in FIG. 29) provides for anapproximately steady fuel-air ratio over a wide range of airflow ratesand venturi temperatures.

-   -   m′.sub.FUEL/m′.sub.AIR.varies.constant

FIG. 29 shows one embodiment of the ejector such as the venturi. In thisembodiment, the size of the opening of the venturi throat 7244determines the amount of suction present at the throat 7244. In aspecific embodiment, the venturi throat is approximately 0.24 inches indiameter. Referring back to FIG. 29, fuel delivery means are coupled tothe venturi 7240. The fuel delivery means may be manifolds, fuel linesor fuel tubes. The fuel delivery means may include other components suchas a fuel restriction 7278, fuel inlet ports 7279 and fuel valves (notshown). Fuel supplied by a pressure regulator 7272 flows through amanifold 7273 and fuel inlet ports 7279 into the relatively lowerpressure in the throat 7244. In one embodiment the fuel inlet ports 7279provide the largest portion of the pressure drop in the fuel deliverymeans. Preferably, making the fuel inlet ports the largest restrictionin the fuel delivery means assures that the restriction occurs at theventuri temperature and maximizes fuel-air mixing by producing thelargest possible fuel plumes. Referring back to FIG. 29, the fuel andair flow into the divergent cone or diffuser 7248 of the venturi, wherestatic pressure is recovered. In the diffuser 7248, the entrained fuelmixes with the air to form an ignitable fuel air mixture in thecombustion chamber 7250. The ignitable fuel-air mixture then enters thecombustion chamber 7250, where the igniter 7260 may ignite the mixture,and the tangential flow induced by a swirler 7230 creates aswirl-stabilized flame. Using an ejector 7240 to draw the gaseous fuelinto the combustion chamber eliminates the need for a high-pressuregaseous fuel pump to deliver the fuel.

In one embodiment, the venturi 7240 is constructed from high temperaturematerials to withstand high temperatures and maintain its structuralintegrity. For the embodiment of FIG. 29, the dimensions of the venturican be approximately 0.9 inches diameter inlet and outlets with anapproximately 0.24 inches diameter throat. The half angles of theconvergent cone and divergent cones can be 21.degree. and 7.degree.respectively and the throat can be 0.25 inches long. In this embodiment,the venturi can be constructed from Inconel 600. Alternatively, otherhigh temperature metals could be used including, but not limited toStainless Steels 310, 316L, 409 and 439, Hastalloy C76, Hastalloy X,Inconel 625 and other super alloys.

In one embodiment, as shown in FIG. 29, a swirler 7230 is locatedupstream of the venturi 7240 and advantageously creates a tangentialflow of air through the venturi. As is well known in the art, thetangential flow from the swirler can create an annular vortex in thecombustion chamber, which stabilizes the flame. Additionally, theswirler 7230 increases the suction pressure at the venturi throat 7244by increasing the local air velocity over the fuel inlet ports 7279.Adding the swirler allows the venturi throat 7244 to be made larger fora given suction pressure. Furthermore, the swirling action induced bythe swirler 7230 can suppress fluctuations in the combustion chamberpressure from propagating upstream to the venturi 7240. Such pressurefluctuations can temporarily slow or stop the flow of fuel gas into theventuri 7240. The swirler 7230 thereby facilitates a steady fuel-airratio in the combustion chamber for steady airflows. The swirler 7230may be a radial swirler.

FIG. 31 depicts an embodiment where an automated controller 7288 adjustsa variable restriction 7292 such as a variable flow valve in the fueldelivery means to hold the exhaust oxygen constant as measured by awide-range lambda sensor or UEGO 7286. In this embodiment, the automatedscheme allows any fuel from biogas to propane to be connected to theburner and the control system can compensate for the changing fueldensity. In this embodiment, the automated controller can restrict thefuel path for dense fuels such as propane and open up the fuel path forlow-density fuels such as methane and biogas. Ignition would beaccomplished by starting the variable restrictor 7292 in the fully openposition, which will produce the richest mixture then closing it untilthe fuel-air mixture is ignited. After ignition, the controller cancontrol the fuel flow to achieve the desired exhaust oxygen level. It isalso envisioned that such an embodiment would allow the fuel air ratioto be adjusted during warm-up to optimize efficiency and burnerstability.

Referring now also to FIGS. 33-34 the gaseous fuel burner 7201 may be ahigh efficiency burner for an external combustion engine such as aStirling cycle engine. In this embodiment, the burner includes an oxygensensor 7286 located in the exhaust stream 7284 and amicroprocessor/controller 7288 to automatically restrict the fuel flowwith the variable restrictor 7292. Additionally, the burner includes ablower controller 7702. The blower controller 7702 may be adjusted bythe microprocessor/controller 7288 to match the Stirling engine poweroutput with the load. In this embodiment, the burner temperature is heldconstant by varying the engine speed and the engine power output isautomatically adjusted by setting the blower speed. Accordingly, in thisembodiment, the burner may burn most gaseous fuels, including fuelswithout constant properties such as biogas.

Referring now also to FIG. 34, fuel may be delivered directly into theventuri at a point proximal to the venturi throat 7244. This embodimentmay include a swirler 7230 to accommodate the fuel delivery means suchas a fuel line or fuel tube. The swirler 7230 may be an axial swirlerpositioned in the venturi 7240 and upstream of the venturi throat 7244.In operation, the delivered fuel is entrained with the motive air toform the fuel-air mixture. In various embodiments, manual or automaticcontrol mechanisms are adaptable to this alternate fuel deliveryembodiment.

Referring back to FIG. 29, the gaseous fuel burner further comprises anigniter 7260 and a flame-monitoring device 7210. Preferably, the igniter7260 is an excitable hot surface igniter that may reach temperaturesgreater than 1150.degree. C. Alternatively, the igniter 7260 may be aceramic hot surface igniter or an excitable glow pin.

With continuing reference to FIG. 29, other embodiments include aflame-monitoring device 7210. The flame-monitoring device 7210 providesa signal in the presence of a flame. For the safe operation of the anyburner, it is important that the fuel be shut-off in the event of aflameout. The monitoring device for flame sensing is the flamerectification method using a control circuit and a flame rod.

Flame rectification, well known in the art, is one flame sensingapproach for the small, high efficiency gas burners. The device uses asingle flame rod to detect the flame. The flame rod is relativelysmaller than the grounded heater head and it is positioned within thecombustion flame. In this flame rectification embodiment, the controlunit electronics are manufactured by Kidde-Fenwal, Inc., and the flamerod is commercially available from International Ceramics and HeatingSystems

Preferably, the flame-monitoring device uses the hot surface igniter asthe flame rod. Alternatively, the flame-monitoring device may be eitherremote from the hot surface igniter, or packaged with the igniter as asingle unit.

Alternatively, an optical sensor may be used to detect the presence of aflame. A preferred sensor is an ultraviolet sensor with a clear view ofthe flame brush through an ultraviolet transparent glass and a sighttube.

It is to be understood that the various fuel burner embodimentsdescribed herein may be adapted to function in a multiple burnerconfiguration as disclosed in U.S. patent application Ser. No.13/447,990, filed Apr. 16, 2012, and incorporated herein.

Some embodiments my control or modulate the flow of gaseous fuel to theburner with reciprocating pumps as disclosed in U.S. patent applicationSer. No. 13/447,990, filed Apr. 16, 2012, and incorporated herein.

Referring now to FIG. 35, a cross section of an engine 9200 is shown.The engine 9200 is similar to the one described above with respect toFIG. 4, however, includes another embodiment of the rocking drivemechanism. The engine 9200 shown in FIG. 35 includes a rocking drivemechanism including link rods 9210, 9210, a rocking beam 9214, a rockingpivot 9224, a connecting rod 9216, a connecting pivot 9218, end pivots9220, 9222, and a crankpin 9226. Although this engine 9200 is an exampleof another embodiment of the rocking drive mechanism as discussed above,the components function in a similar fashion however, this embodimentincludes a number of additional benefits.

The configuration of the connecting rod 9216, rocking beam 9214, andconnecting pivot 9218 limit the loads on the connecting rod. Thisconfiguration additionally allows for the use of larger bearings,including standard sized tri-metal bearings. Additionally, the increaseddistance between the rocking pivot 9224 and the connecting pivot 9218increases the mechanical advantage of the rocking pivot 9224, thusreducing the loads on the connecting rod bearings

In this embodiment of the engine 9200, the side loads on the link rods9210, 9210 has been increased. However, as discussed above, the engine9200 is an oil lubricated engine, thus, concern with limiting the sideloads on the link rods 9210, 9210 has been reduced. Thus, in theembodiment shown in, for example, FIG. 4, the link rods are longer andthe loads on the connecting rod are higher. In the embodiment shown inFIG. 35, the link rods 9210, 9210 are shorter and the load on theconnecting rod 9216 is decreased.

In some embodiments, the oil pump is a Gerotor pump driven by thecrankshaft through a spline connection. In some embodiments, the oilpump is driven by the crankshaft by a gear.

Referring now to FIG. 36A, an embodiment of a Stirling cycle machine isshown in cross-section and designated generally by numeral 9600. Whilethe Stirling cycle machine 9600 will be described generally withreference to the embodiment shown in FIGS. 36A and 36B, it is to beunderstood that many types of machines and engines, including but notlimited to refrigerators and compressors may similarly benefit fromvarious embodiments and improvements which are described herein,including but not limited to, external combustion engines and internalcombustion engines. In particular, the present embodiment of theStirling cycle machine is directed to improving the efficiency andoperation of a 10 Kilowatt (Kw) Stirling cycle machine, although anyother power output level are certainly contemplated and encompassedwithin the following disclosure of a machine or engine, that achieveshigh efficiency, long durability and low cost targets based onsimultaneously utilizing optimized mechanical and operational controlsystems with existing Stirling cycle machine platforms.

The engine 9600 shown in cross-section in FIG. 36A includes generally acrankcase 9610 housing the drive components of the engine and a workspace 9620 containing the working gas and/or fluid and gas and/or fluidcompression and expansion related components. Inside the crankcase 9610is an embodiment of a rocking beam drive mechanism 9601 (the term“rocking beam drive” is used synonymously with the term “rocking beamdrive mechanism”) for an engine, such as a Stirling engine, havinglinearly reciprocating pistons 9602 and 9604 housed within cylinders9606 and 9608, respectively. As discussed previously, rocking beam drive9601 converts linear motions of pistons 9602 and 9604 into the rotarymotion of a crankshaft 9614. Rocking beam drive 9601 has a rocking beam9616, rocker pivot 9618, a first coupling assembly 9619, and a secondcoupling assembly 9621. Pistons 9602 and 9604 are coupled to rockingbeam drive 9601, respectively, via first coupling assembly 9619 andsecond coupling assembly 9621. The rocking beam drive 9601 is coupled toand drives crankshaft 9614 via a connecting rod 9622.

This embodiment shown in FIGS. 36A and 36B is an inverted rocking beamdesign similar to that disclosed in FIG. 35 and incorporates the sameadvantages and benefits as discussed therein. An important advantage ofthe inverted rocking beam arrangement having the crankshaft 9614 locatedrelatively below the rocking beam mechanism 9601 of the machine ensuresthat the structural arrangement and alignment of the piston rods 9624and cross-head coupling means 9634 which connect the piston rods 9624 tothe rocking beam drive 9601 do not have to account for the size of thecrankshaft 9614 and related components. This arrangement facilitates alarger load carrying capacity conrod bearing 9615 on the connecting rod9622, better mechanical advantage developed by the rocking beam 9616 andspace for such larger conrod bearings 9615. The arrangement alsorelieves space constraints of the pistons 9602 and 9604, piston rods9624 and cylinders 9606 and 9608 which can occur with the crankshaftlocated above the rocking beam drive and between the piston shafts 9624.With the rocking beam 9616 now located above the crankshaft 9614, thereare no longer space restrictions around the crankshaft rocking beam 9616and a larger wrist pin bearing 9628 can be provided to better supportthe connecting rod 9622 and rocking beam connection 9601.

Also, with the inverted design, the rocking beam 9616 can be designed toreduce the load on the connecting rod 9622 wrist pin bearing 9628 andconrod bearing 9615 by adjusting the lever arm ratio A and B seen inFIGS. 37A and 37B between the rocking shaft pivot 9718, the connectingrod wrist pin bearing 9728 and between the rocker shaft pivot 9718 andthe lever arm of the piston acting at 9729. For example, as seen in FIG.37A, the bearing load on the connection rod 9728 is greater where theconrod bearing ratio A/B is 1.6 relative to the piston connection. InFIG. 37B, the rocking beam 9616 is shown having a 1.0 ratio whichessentially equates the distances of the two connection points 9728′ and9729 about the rocking shaft pivot 9718 and therefore correspondinglybalances the load on the crankshaft 9614 to be the same or similar tothat developed by the piston shaft 9624 and significantly lower than theload transmitted with the bearing ratio A/B of 1.6. It is to beappreciated that other embodiments of a rocking beam drive besides theinverted rocking beam drive may also incorporate the benefits of thedisclosed rocking beam 9616 as well.

Referring also to FIGS. 36A-36E, the alignment of the pistons 9602 and9604, piston rods 9624 and cylinders 9606 and 9608 in conjunction withthe crankcase 9610 is of critical importance for power transmissionthrough the pistons 9602 and 9604 and piston rods 9624, providing forreduced wear on the piston rings and to the dynamic alignment andreciprocating nature of the piston rods and high pressure piston rodbearings 9630. The crankcase 9610 contains most of the rocking beamdrive 9601 and is positioned below the cylinder housing 9631. Thecrankcase 9610 defines a space to permit operation of rocking beam drive9601 having the crankshaft 9614 located below rocking beam 9616, aconnecting rod 9622, and first and second coupling assemblies 9619 and9621. Pistons 9602 and 9604 reciprocate in respective cylinders 9606 and9608 as also shown in FIG. 36 and cylinders 9606 and 9608 extend abovecrankcase 9610, through the cylinder housing 9631, and into the heaterheads.

The cross-heads 9634 and cross-head bores 9635 have a tolerance that isdifficult to align with that of the mating cylinder liners in thecylinder housing 9631 during what is referred to as “stack up” i.e. thejoining of the separate parts of the vessel, and therefore anydifference in concentricity between these two elements when they areassembled together, creates a potential for misalignment, wherepotentially, the piston rod 9624 could sit askew, or at an angle andtherefore the piston may reciprocate non-coaxially to the cross-heads9634. The cylinder liner bores 9606, 9608, the cylinder gland locatingdiameter and the cross-head locating diameter of the cross-head bore9635 all must be in alignment. To alleviate this issue and potential formisalignment, all three of these diameters are bored together in thesame set-up and essentially simultaneously in the same operationresulting in very close tolerances of the diameters and theconcentricity of these elements is maintained as closely as possiblebased on machining tolerances. These elements may also be manufacturedand bored in other ways as well including but not limited to withalignment jigs and separate boring process that can produce therequisite tolerances to ensure that any angular deviation of the pistonis maintained within an acceptable range.

Also as shown in another embodiment in FIG. 36B, to improve theconcentricity of the piston and piston rod 9624, each piston rod 9624 isprovided with a tapered end 9625 at each end of the rod 9624 to wedgethe first end of the piston rod into the cross-head 9634. The taperedend 9625 facilitates the location, resting and clamping (L,R,C) betweenall elements for a proper location of the piston rod 9624 with thediameter doing the locating, the taper 9625 doing the resting, and a nut9633 of the end doing the clamping. In the wedge connection provided bythe taper, the wedge can lock itself in place because of the loadsdeveloped by the piston. The wedge or tapers on the ends of the pistonrods are essentially jammed more and more firmly into the crossheads9634 at the lower end of the piston rods 9624 and correspondingly intothe piston at the upper end of the piston rod. A nut 9633 may be used tofacilitate the connection with the cross-head 9634 in case the rod comesloose, but in almost every case the wedge will maintain the appropriateconnection of the piston rod 9624 to the piston above, and thecross-head 9634 below.

To facilitate the assembly of the tapered piston rods 9624 where thetaper is essentially a reduction in diameter of the ends of the pistonrod 9624 along a portion of the piston rod, the piston 9602, 9604 ismanufactured from two separate parts, a piston base 9643, and a pistonshell 9645 better shown in FIGS. 36C, D and E. The piston base and shellcan be matingly threaded where the piston base 9643 defines a threadedinner diameter surface wall 9647 corresponding to a threaded outersurface wall 9649 of the piston shell 9645. Other connectionarrangements between the base and shell are possible as well tofacilitate the connection of the two piston elements. The piston base9643 is provided with a receiving bore 9651 which may be a constantdiameter bore, or a tapered bore to receive the tapered end of thepiston rod 9624. To assemble these elements, the tapered piston rod 9624is inserted into the piston base 9643, clamped in place with a desiredpre-load, and then the shell 9645 is threaded onto the base 9643 tocomplete the assembly. The reason for the two-part piston is that toappropriately clamp and pre-load the piston rod 9624 to the base 9643,the assembly process necessitates access to the inside of the piston,and hence the two-part shell and based design facilitates the clampingprocess. Other manufacturing techniques may also be used toappropriately attach the tapered piston rod 9624 and piston 9602, 9604without the necessity for a two-part piston as described above.

Another important aspect of the present embodiment is an increasedvolume of the combustion space in the heater head. To provide morevolume for the combustion of the burner to take place and heat thetubes, an upper most portion 9655 of the cylinders 9606, 9608 isprovided with a larger diameter than the remainder lower portion of thecylinders, giving the cylinders 9606, 9608 to some extent amushroom-shaped profile. The benefit of this includes but is not limitedto the ability to move the heater tubes 9659 farther out from an axialcenter of the cylinders 9606, 9608, thereby increasing the diameter andcombustion volume above the cylinder inside the heater tubes 9659 and/orto accommodate a larger diameter tube to handle more working gas andfluid through the heater tubes 9659.

In another embodiment shown in FIG. 38B, a Gerotor displacement pumpingunit is driven by the crankshaft 9814. The Gerotor pump uses an innerrotor 9844 having one less gear tooth 9846 than the surrounding outerrotor 9848. During part of the rotation cycle, the area between theinner and outer rotor increases, creating a vacuum that draws fluidthrough an intake. The area between the rotors then decreases, causingcompression allowing oil to be pumped out to the mechanical parts of theengine. The Gerotor pump is driven coaxially and directly from thecrankshaft 9814 without the transmission losses of the helical drivegear, making the engine construction and assembly more efficient andless expensive than the construction and components of the helical drivegear to the gear pump. The construction and assembly is easier becausethe Gerotor pump is directly driven by the crankshaft 9814, whereasthere are significant mechanical losses associated with the previousgear pump. In other embodiments different pumps besides a Gerotor pumpmay be used, which include but are not limited to, gear pumps, pistonpumps, rotary gear pumps, hydraulic pumps and diaphragm pumps forexample and that other embodiments of a rocking beam drive besides theinverted rocking beam drive may incorporate the benefits of the Gerotorpump or similar direct drive pump.

High Pressure Rod Seals

The present embodiment of the Stirling cycle engine maintains theworking space 9620 and the working gas and/or fluid at a relatively highpressure, generally in the range of 1200-1800 psi, and more preferablyabout 1500 psi. It is of course necessary to ensure that the working gasand/or fluid is essentially sealed in the working space 9620 so that itdoes not escape into the crankcase 9610 and the environment. A criticalplace for such leakage of working fluid to occur is around the pistonrods 9624, which extend and reciprocate between the working space 9620and the crankcase 9610. To minimize such leakage, a high pressure pistonrod seal 9630 is provided below the respective cylinders 9606 and 9608and between the working space 9620 and the crankcase 9610.

With a significantly higher pressure in the working space 9620 relativeto the crankcase 9610, a certain amount of working gas is anticipated toleak through the high pressure rod seals 9630. However, it is imperativeto minimize the leakage without significantly affecting thereciprocating efficiency of the pistons and the engine. Also, as will bediscussed in further detail below, an airlock and working fluidrecapture system may be used in conjunction with the high pressure sealsto capture certain amounts of such leaking working gas and/or fluid. Anyworking gas which leaks into the air lock between the working space 9620and the crankcase 9610 can be drawn into an accumulator and suppliedback into the workspace when necessary. Before more completelydiscussing such an airlock and recapture of working fluid, the presentdiscussion is focused on the use of the high pressure rod seals 9630between the working space 9620 and the crankcase 9610 to ensure the mosteffective working fluid pressure and gas containment.

A mechanical embodiment of the high pressure rod seal 9930 is in FIG.39. It should be understood that such a rod seal is intended to beutilized not only in the Stirling engine embodiments described hereinbut also in other engines or mechanisms with similar reciprocatingpistons.

In a mechanical embodiment of the high pressure piston rod seal 9930′,shown in better detail in FIG. 39, a substantially symmetricalhemispherical shaped piston sleeve 9960 is supported by an upper sealsupport 9965 and a lower seal support 9966 inside a seal cavity definedinside a seal housing 9951. The symmetry of this hemispherical shapedpiston sleeve 9960 provides more consistent wear across the length ofthe sleeve 9960 as compared to the wedge rod seal 9930 described abovewhich focuses the radial wear at one end of the sleeve. Thehemispherical surface 9963 of the piston sleeve 9960 bears on an innerrespective bearing surface of each of the upper and lower seal supports9965, 9966. A wear support clamp 9967 is provided axially disposed abovethe upper seal support 9965 which forces the upper and lower sealsupports 9965, 9966 into biased contact with the piston sleeve 9960. Agap G may be provided between the upper and lower seal supports 9965,9966 to accommodate any wear that may occur on abutting surfaces in theseal. As wear occurs, the abutting surfaces in the seal may be reducedso that as the sleeve bearing wears, the upper and lower seal supports9965, 9966 are biased towards one another by the support clamp 9967. Thegap G permits the upper and lower seal supports 9965, 9966 to movecloser to one another as the seal wears without interfering with oneanother and so maintaining contact with the hemispherical shaped outersurface 9963 of the piston sleeve 9960.

A still further embodiment of a high pressure rod seal shown in FIG. 40Aincludes a spring energized lip seal 10003 generally comprising a sealjacket, made from PTFE or graphite for example, and a spring (not shown)circumferentially secured within a groove or between lips 10007 of theseal 10003. When the spring energized lip seal 100003 is seated in thehousing, the spring lip seal 10003 is under compression, forcing thejacket lips 10007 against the respective adjacent walls of the sealblock 10011 and the surface of the reciprocating piston 10024, therebycreating a leak free seal. The lip seal 10003 provides permanentresilience to the seal jacket 10005 and compensates for jacket wear andhardware misalignment or eccentricity. System pressure also assists inenergizing the seal jacket 10005. Spring loading assisted by systempressure provides effective sealing at both high and low pressures.Spring energized lip seals are highly durable and designed for static,rotary and reciprocating applications in temperatures from cryogenic to+600F as well as pressures from vacuum to 25,000 psi, and to survivemost corrosive environments.

A spring cup retaining cylinder 10008 is set around the piston rod 10024and supported on a lower collar 10006. The retaining cylinder 10008maintains a circumferential space about the piston rod 10024 in whichthe lip seal 10003 is maintained. The lip seal 10003 can be a PTFE andgraphite ring supported around an outer circumference by the retainingcylinder 10008 and frictionally slidably engages the piston rod 10024 tocreate the high pressure spring energized lip seal. The spring (notshown) inside the lip seal 10003, along with the higher pressure of theworking space, forces the lip seal 10003 against the respective pistonrod 10024 and retaining cylinder wall, and also maintains the lip seal10003 set down in the retaining cylinder 10008 generally against thelower collar 10006.

A hydraulic embodiment of a high pressure piston rod seal can facilitatean efficient and long term seal between the working space and theairlock. FIG. 40B discloses a hydraulic high pressure piston rod seal10021 set inside the rod seal cavity of a test rig. A rod seal sleeve10023 circumferentially encompasses the piston rod 10024 and defines apressure space 10025 between a wall of the test rig and an outer surfaceof the rod seal sleeve 10023. A hydraulic fluid pressure line 10027communicates with pressure space 10025 to provide the appropriate fluidpressure to maintain the rod seal sleeve 10023 in sealing engagementwith the piston rod 10024. A sensor (not shown), such as apiezo-electric pressure sensor, can be provided in the pressure space10025 and on the rod seal sleeve 10023 to ensure that the appropriatepressure and flexure is actuating the rod seal sleeve 10023 andproviding the appropriate sealing pressure against the piston rod 10024.The inner surface of the rod seal sleeve 10023 slidably engages alongthe piston rod 10024 as the rod reciprocates and the rod seal sleeve10023 is influenced radially inwards by the hydraulic pressure fluid inthe pressure space 10025. As the rod seal sleeve 10023 wears, thehydraulic fluid pressure in the pressure space 10025 can be increased toensure that the rod seal sleeve 10023 is motivated radially towards thepiston rod 10024 to maintain slidable engagement with the piston rod.

It is to be appreciated that the above disclosed embodiment of highpressure rod seals are intended only as examples and that the machinesdescribed herein are not limited to these examples, and that otherembodiments of high pressure rod seals may also be used to ensure thatthe high pressures used in Stirling engines, or any other engine forthat matter, are maintained in the appropriate working space, crankcaseand other engine compartments as necessary.

Rolling Diaphragm Seal

Turning to FIGS. 41A and 41B, and referring back also to FIGS. 10A-D, incertain embodiments a rolling diaphragm 10190 is used in conjunctionwith the piston rods 10124 to prevent the escape of lubricating fluidfrom the crankcase 10110 up past the rods 10124 and into the workingspace 10120 and regenerator. If the lubricating fluid necessary for therocking drive can bypass the piston rod seals, it can potentially damagethe working space, clog the regenerator and contaminate the workingfluid or gas of the engine in the cylinders.

To facilitate the appropriate rolling and flexing of the diaphragm10190, a pressure differential is maintained across the rollingdiaphragm 10190 so that preferably the pressure above the diaphragm10190 is slightly greater than the pressure in the crankcase. The sealis thus essentially inflated into the crankcase, which facilitates thediaphragm 10190 maintaining its desired form as it rolls and flexes withthe reciprocating piston rod 10124. This alleviates stresses on thecircumferential sealing points so the seal is not compromised. It isgenerally necessary to place a differential of approximately 15 PSIacross the diaphragm 10124 to properly inflate the seal so that itconforms to the shape of the bottom seal piston 10195 as it moves withthe piston rod 10124. It is to be appreciated that the pressuredifferential maintained across the rolling diaphragm 10190 is notlimited to 15 PSI. Rolling diaphragms made of stronger materials orhaving a particular shape may be able to sustain a higher differentialor operate at a lower differential as the case may be. In embodiments ofthe stirling cycle engine where the working space 10120 is at arelatively high pressure 1500 PSI-1800 PSI, the crankcase 10110 must becharged with a pressure for instance of 1485 PSI, which is approximately10-15 PSI less than that of the working space at 1500 PSI. Although itis possible to regulate these larger pressures to maintain the 10-15 PSIdifference across the diaphragm, it is difficult and adds to thecomplexity of the machine.

The rolling diaphragm 10190 may be manufactured by injection molding orhot compression molding. In hot compression molding of the rollingdiaphragm 10190, it can be more difficult to control material propertiesbut injection molded diaphragms have shown in testing a bettertransition of dynamic stresses across the profile of the rollingdiaphragm 10190 as it transitions and rolls with the reciprocation ofthe piston rod 10124. Testing on the materials used to fabricate therolling diaphragm 10190 indicate chopped fiber is most successful forexample but not limited to, nitrile with Kevlar fiber or Fab-Air®.

FIGS. 41A and 41B disclose an embodiment of the rolling seal ordiaphragm 10190 having a profile which facilitates the dynamic rollingtranslation of the diaphragm. As discussed in previously herein, andincorporated herein by reference in its entirety in the presentdiscussion, the pressure differential that is placed across the sealallows the seal to act dynamically to ensure that the rolling diaphragm10190 maintains its form throughout its dynamic range of motion. Aspreviously discussed, the pressure differential causes the rollingdiaphragm 10190 to conform to the shape of the bottom seal piston 1310with reference to FIG. 10A as it moves with the piston rod 10124, andprevents separation of the diaphragm 10190 from the surface of thepiston rod 10324 during operation. It is desirable to lower the amountof inflation of the rolling diaphragm 10190 without the diaphragmbuckling or separating, i.e., deviating from a consistent dynamic axialand radial rolling of the diaphragm 10190 along the diaphragm profilewith the axial reciprocation of the piston rod 10124. As discussedabove, the inflation of the diaphragm is provided by the pressuredifferential across the rolling diaphragm 10190. To accomplish this, ithas been found that particular structural profiles facilitate theconservation of material and consistent rolling of the diaphragm.

The cross-section in FIG. 41A-B shows a profile view of the molded formof the diaphragm of the present embodiment about a diaphragm axis L. Forpurposes of describing the diaphragm structure the inner edge 10192 asbeing the top 10194 of the diaphragm and the outer edge 10193 is thebottom of the diaphragm as shown in the figures. The diaphragm has alateral wall 10190 extending axially and radially relative to axis Lfrom the inner edge 10192 to the outer edge 10193; the lateral wall iscomposed of several sections. A top fillet section 10198 turns thematerial approximately 90 degrees from the top of the diaphragm 10190 asshown, to a sidewall section 10196 substantially parallel to the pistonrod 10124 and axis L. The sidewall section 10196 in turn then turnstowards the outer edge 10193. Before reaching the outer edge 10193, thesidewall section merges contiguously into a chamfer section 10199, whichwhile still depending axially from the sidewall section 10196, extendsfrom the sidewall 10190 in a greater radial degree relative to axis L toconnect with the outer edge of the diaphragm 10193. The sidewall section10196 may be parallel to the axis L or may also have a radial componentwhich slightly angles the sidewall section 10196 radially away from theaxis L. In either case the chamfer section 10199 extends to a greaterradial degree from axis L than the sidewall section 10196. A bottomfillet 10197 connects to the outer edge 10193 defining the bottom of thediaphragm as drawn. The outer edge 10193 like the inner edge 10192 isprovided with a thickened circumferential lip, which can be securedinside a matching groove formed in the vessel joint.

The cross-section shown in FIGS. 42A and 42B is a profile view of themolded form of another embodiment of the rolling diaphragm 10290 of thepresent embodiment about a diaphragm axis L. Like reference numbers forthis embodiment correspond to the same or similar elements in theprevious rolling diaphragm embodiment. For purposes of describing thediaphragm structure, the inner edge 10292 is the top of the diaphragmand the outer edge 10293 is the bottom of the diaphragm 10290 as shownin the figures. The diaphragm 10290 has a lateral wall 10296 extendingaxially and radially relative to axis L from the inner edge 10292 to theouter edge 10293; the lateral wall here is again composed of severalsections. A top fillet 10294 section turns the material approximately 90degrees from the top of the diaphragm 10290 as shown, to the sidewallsection 10296, which depends both axially and radially outwards towardsthe bottom of the diaphragm along the axis L. Before reaching the outeredge 10293, the sidewall section 10296 merges contiguously into a bottomfillet 10299 to extend towards the outer edge 10293 of the bottom of thediaphragm as drawn. An outer lip 10197 similar to the thickenedcircumferential lip 10295 of the inner edge 10192 is provided, which aresecured inside a matching groove formed in the vessel or crankcase jointwhich secures and seals the outer edge 10293 of the diaphragm.

The injection molding of the diaphragm is important because the gatingmethods and other molding techniques, characteristics, methods andspecifications can affect the fiber alignment and molecular alignment ofthe diaphragm material during the molding process. These materialcharacteristics are important because this can affect the hoop stress ofthe diaphragm. For example, if the material is gated at one end andoverruns an opposing end of the mold, the fibers can be aligned in aparticular direction to optimize the hoop strength of the diaphragmwhile potentially enhancing the flexible and rolling characteristics ofthe final diaphragm element.

It is very important in the dynamic rolling actuation of the diaphragms10190, 10290 that no imperfections or particles including fluidparticles such as oil droplets are disposed on the surfaces of thebottom seal piston or on the adjacent vessel wall surrounding the bottomseal piston. Such fluid particles, most likely oil, are detrimental tothe rolling actuation of the diaphragm against the respective crankcasesurfaces, because they cause stress points on the diaphragm.

Turning to FIG. 43 another embodiment of the rolling diaphragm includesa first and second rolling diaphragm 10391, 10393 to make what isessentially a double bellows system 10392. A double bellows system 10392can facilitate the elimination of the 10-15 PSI pressure differentialbetween working space and airlock and/or crankcase by providing theappropriate expansion pressure between the double bellows themselves.The double bellows include first and second rolling diaphragms 10391,10393 which are oppositely and axially aligned along the piston rod, anddefine a space therebetween with a light oil contained between thediaphragms and pressure charged between the double bellows. Theincompressible oil prestresses the diaphragms and facilitates theconsistent rolling of the diaphragm as the piston rod 10324 reciprocatesalong its axis.

Airlock and Working Fluid Recapture System

The power, life and value of a Stirling engine can be maximized bybuilding an oil lubricated drive and sealing the work-space from theoil. Oil lubricated drives allow high powers and are inexpensivecompared to drives based on rolling elements. It is essential to isolatethe oil in the drive from the workspace. Even oil mist will migrate tothe hot end of the working space, where the oil will breakdown and theresulting carbon will clog the heat exchanger. Flexible membranes orbellows such as the rolling diaphragms discussed above that attach tothe moving piston rod and the structure provide an oil and gas tightseal between the oil filled crankcase and the workspace, ensuring thatthe lubricant is maintained in the crankcase. In order to function forthousands and millions of cycles, a small pressure difference must bemaintained across the bellows.

An important aspect of the rolling diaphragm and oil lubricatedcrankcase relates to the use of an airlock 10401 and an airlock pressureregulation system 10411 as shown in FIGS. 44A and 44B. The airlockpressure regulation system 10411 provides the benefit of ensuringworking gas escaping from the working space 10403 is returned to theworking space, provided that the working gas does not leak into theenvironment or atmosphere, which would require replenishment of theworking gas, and that an appropriate pressure differential is maintainedacross the rolling diaphragms as described above. The airlock pressureregulation system 10411 permits an easily serviceable bottom end i.e.crankcase 10410 if, as in the embodiment disclosed in FIG. 48, thecrankcase is intended to be maintained essentially at atmosphericpressure.

As shown in FIG. 44A relating to a pressurized crankcase 10410 atapproximately 1485 PSI, in order to maintain an appropriate workingspace pressure and airlock pressure regulation, an airlock space 10401is provided between the working space 10403 and the crankcase 10410 at apressure of, for example 1500 PSI, so that the substantially greaterpressures in the working space 10403 should not significantly influencethe air lock space 10401 and any pressure and working gas leaking fromthe working space 10403 into the air lock can be captured andaccumulated as described below with respect to the airlock pressureregulator and returned to the airlock and working space and not merelyescape into the crankcase and environment.

It is to be understood that airlock space 10401 is intended to maintaina constant volume and pressure necessary to create the pressuredifferential necessary for the function of rolling diaphragm 10490 aspreviously described. In the present embodiment the airlock 10401 may ormay not be sealed off from working space 10403 with high pressure rodseals 10430. In any case, the pressure of airlock space is desired to bemaintained at essentially 1500 PSI and equal to the mean pressure ofworking space 10403. The pressure in the working space 10403 may vary atleast +/−300 PSI so the intention of the airlock space 10401 is toinsulate the diaphragms from such fluctuations and maintain itself ataround the necessary pressure, by way of example here 1500 PSI, relativeto the 1485 PSI charged in the crankcase 10410. To facilitate theequalization of pressures between the working space 10403 and theairlock space 10401, a small opening or pressure equalization orifice10404 communicates between the working space 10403 and the airlock space10401. The crankcase 10410 must be charged to 1485 PSI, and bemaintained at approximately 15 PSI less than the airlock space 10401 sothat the appropriate pressure is applied to the rolling diaphragm 10490to ensure the proper dynamic movement of the diaphragm.

In this pressurized crankcase 10410 embodiment an airlock pressureregulator 10411, a pump 10412 and relief valve system is providedbetween the crankcase 10410 and the air lock space 10401 to maintain theexemplary 15 PSI pressure differential therebetween. Other predeterminedpressure differentials may also be maintained depending on the diaphragmmaterial and the design of the entire airlock pressure regulationsystem. In its most general form, an uptake line 10416 communicates fromthe pressurized crankcase 10410 to a a filter 10418, a pump 10412(having a check valve on its outlet), and a pressure regulator 10413 inparallel with the pump 10412 and filter 10418 for returning pressurizedworking gas back to the air lock 10401 and so maintains the pressuredifferential between the airlock space 10401 and the crankcase 10410 andconsequently across the rolling diaphragm 10490. This airlock pressureregulator system 10411 is described more completely with respect to FIG.44B.

The airlock pressure regulator 10411 regulates the pressure differencebetween the airlock 10401 and the crankcase 10410. When the engine isturning, the airlock pressure regulator 10411 keeps the airlock pressurepreferably 10 to 14 PSI above the crankcase pressure although a range of5 to 20 PSI is possible and other pressure differentials can beaccomplished by the regulator as well. When the engine is off, theairlock pressure regulator 10411 keeps the airlock pressure preferablyless than 15 PSI above the crankcase pressure and not more than 5 PSIbelow crankcase pressure. It is permissible to have a greaterfluctuation of pressure differential when the engine is off since thereis little or no dynamic forces being applied to the rolling diaphragms10490 via moving pistons.

The airlock pressure regulator 10411 performs several importantfunctions. The airlock pressure regulator 10411 uses a pump 10412 tomove pressurized gas from the lower pressure crankcase 10410 into theairlock 10401, thereby maintaining the airlock 10401 at a higherpressure. The airlock pressure regulator 10411 relieves excess pressurebetween the airlock 10401 and crankcase 10410 volumes. A bidirectionalregulator 10413 vents some of the airlock gas into the crankcase 10410,when the airlock pressure is preferably 15 PSI above the crankcase andvents in the opposite direction, venting gas from the crankcase 10410 toairlock 10401, when the airlock pressure is more than 5 PSI below thecrankcase pressure. Also, a filter 10418 in the airlock pressureregulator 10411 filters out the oil from the crankcase gas before itenters the airlock volume.

The components of the preferred embodiment are the mechanical pump10412, the bidirectional pressure regulator 10413, an oil filter 10418,a pump pressure switch 10417 to control the pump 10412 and a controllerpressure switch 10419 to signal the engine controller C. An example ofthe mechanical pump is the Linear AC 0410A pump by Medo. Other pumpscould certainly be used as well. The important qualities of the pump arethe ability to operate in a high pressure inert environment, long life,no maintenance and quiet. Solberg Mfg. produces a line of oil-misteliminators, i.e. filters, that are compact, effective and can holdenough oil for several thousand hours of operation. In a preferredembodiment, the bidirectional regulator 10413 allows pressure flow whenthe design pressure difference has been exceeded in either direction.Pump pressure switch 10417 operates the pump when the pressuredifference between the airlock 10401 and the crankcase 10410 ispreferably less than 10 PSI for example. Pump pressure switch 10417includes a predetermined dead band, or range, that keeps the pump 10412on until the airlock pressure is for example 14 PSI above the crankcasepressure. Controller pressure switch 10419 signals to the controller Cthat the airlock pressure is at least 5 PSI, for example, above thecrankcase pressure. This insures that the engine will not turn until theairlock pressure is sufficiently greater than the crankcase pressure.The rolling diaphragms 10490 could tear if moved without such pressuredifference across them. A fill source 10414 may be connected with theairlock to replenish the pressurized vessel charging and workinggas/fluid if necessary.

FIG. 45 is a specific embodiment of the bidirectional regulator 10413showing the pump 10412, oil filter 10418 and a spool valve 10441 whichoperates between an airlock port 10449, a crankcase port 10451 and apump port 10453 according to the pressure differentials between thecrankcase pressure and airlock pressure. In this case, alternative tothe pressure switches 10417, 10419 described above a proximity sensor10425 for determining location of the spool 10441 via a target magnet10426 is used to control the pump 10412 and if necessary to signal theengine controller C. The spool valve 10441 is biased by a primary spring10443 against the airlock over-pressure and an underpressure reliefvalve 10445 is biased by an inner spool spring 10447. Observing FIGS.46A-A6E the spool is shown in certain positions: in (1) is shown thespool influenced open by the spring where the airlock pressure is low sothat the airlock port 10449 now communicates to pump port 10453 toreceive pressurized gas from the pump 10412, in (2) the spool 10441 isshown where the airlock pressure is within normal limits so the airlockport 10449 is closed by spool 10441 and the spool is still displacedenough according to the proximity sensor 10425 to cause operation of thepump 10412, even without flow from the pump to the airlock. In (3) thespool 10441 is shown where the airlock pressure is again within normallimits so the airlock port 10449 is closed by spool 10441 and the spoolis now displaced so that the proximity sensor 10425 does not turn on thepump. Either one or two proximity sensors 10425 are shown in FIGS. 45-D,however any desired number and type of proximity sensors may be used innormal operation in other embodiments. In (4) the spool 10441 is shownwith the airlock pressure is high so that the airlock port 10449 isconnected to crankcase port 10451 and pump is disabled while airlockpressure is reduced. (5) is a case where the engine is shut down sothere is no power to the pump and the airlock pressure is extremely lowand to keep the diaphragms from being damaged, the airlock port 10449 isconnected directly to the crankcase through the underpressure reliefvalve 10445 which opens to provide direct pressure relief through thespool 10441 so that the crankcase pressure and airlock pressure are atleast equalized.

In another embodiment of the pressure regulator 10401 shown in FIG. 47,the bidirectional regulator 10413 is replaced with a back pressureregulator 10431 which provides for one way pressure flow from theairlock 10401 into the crankcase 10410 should the pressure differentialexceed for instance 15 PSI. To accommodate flow in the other directionfrom the crankcase to the airlock, a check valve, or pair of checkvalves 10433, 10435 can be provided in a separate path. This ensuresthat the crankcase will not be pressurized higher than the airlock. InFIG. 48, the crankcase 10510 is intended to be maintained at atmosphericpressure. This is a critical improvement as it provides for a moreeasily serviceable lower unit on the vessel without the need to rechargethe crankcase 10510 should work need to be done inside the crankcase10510 and also provides that a significantly lighter crankcase housingis necessary to contain the drive components. In this embodiment of theairlock pressure regulator system 10511 the airlock space 10501 ismaintained essentially at atmospheric plus 15 PSI and therefore anypressurized working gas which escapes from the working space 10503 intothe airlock 10501 needs to be removed from the airlock 10501 andreturned to the working space 10503. To accomplish this, in its simplestform a first relief valve 10520 means is provided in an uptake line10522 communicating with the airlock space 10501 so that any pressuregreater than 15 PSI is relieved from the airlock 10501 and passed via apump 10512 into an accumulator 10523 outside the working space 10503,airlock space 10501 and crankcase 10510. From the accumulator 10523 areturn line 10525 includes a second relief valve 10521 means which opensto permit recharging of the working space 10503 with pressurized gasfrom the accumulator 10523 should the pressurized working gas in theworking space 10503 fall below 1500 PSI. It is to be appreciated thatthe balancing of this pressurized system may include other pressureconsiderations across the first and second relief valves 10520, 10521,particularly with respect to the variation which can occur in theworking space 10503 where the pressure can swing plus or minus 300 PSIduring the Stirling cycle itself.

When the engine is running, a mechanical pump 10612 defined by a cavity10608 in the piston rod may be utilized to reduce the load and work doneby the above described airlock pressure regulator system. As seen inFIG. 49, the mechanical pump 10612 of the piston rod 10624 is added tothe airlock pressure regulation system to reduce the load on theelectrical system during operation. A check valve 10605 receivescrankcase pressure through an intermediate passage 10607 from thecrankcase. The check valve 10605 opens when the airlock pressure hasdropped too low relative to the crankcase pressure and pressurized gasfrom the crank case is drawn into the piston cavity 10608 as the pistonrod 10624 reciprocates. The piston rod cavity 10608 is defined by areduced diameter portion of the piston rod which essentially defines themechanical pump 10612 itself. As the piston rod 10624 reciprocates thepiston rod cavity 10608 is reduced in size as shown by the right-handpiston, pumping the pressurized gas into the airlock space 10609. Inthis way during engine operation the airlock can be efficientlyreplenished with sufficient pressurized gas should its pressure drop toolow. An outlet check valve 10611 is provided between the airlock and thecrankcase so that pressure in the airlock which exceeds the desireddifferential can be reduced from the airlock space 10609 into thecrankcase. The mechanical pump 10612 defined by movement of the pistonrod 10624 does not operate at engine startup because there is nomechanical operation of the engine, however the airlock pressureregulator system must be operational during startup operations.

Cooler Liner Diameter Reduction

As explained previously with respect to FIGS. 1, 4-9C and 19, the heatertubes communicate with a heat exchanger which circumferentiallysurrounds each cylinder. The heat exchanger of the present embodimentdescribed in FIG. 50A provides cooling for the working gas/fluid duringthe appropriate portion of the stirling cycle. The heat exchanger 10705is supplied with cooling water through coolant tubing which communicateswith a heat sink such as the environment via a radiator (not shown).Generally, the coolant water picks up heat through the heat exchanger inthe vessel from the hot working gas, and the coolant water then ispumped to the radiator where the heat is transferred to the environment.

The heat exchanger 10705 shown in FIG. 50A surrounding each respectivecylinder is provided with a water jacket sleeve 10704 having an innersurface defining a channel to allow passage of the cooling water throughan interfacial area 10706 between the inner surface of the water jacketsleeve 10704 and an outer surface of a cooler liner 10702. The coolerliner 10702 also has an inner surface 10708 which directs the flow ofhot working gas along the inner surface to facilitate the transfer ofthe heat through the cooler liner 10702 to the coolant water in theinterfacial area. A goal of the described structure is to increase theheat transfer surfaces within the interfacial area 10706 for absorbingheat from the hot working gas and the heat exchanger so as to improveheat transfer between the working gas and the coolant water.

The water jacket sleeve 10704 surrounds the cooler liner 10702 and formsthe heat exchanger 10705 which cools the working fluid during theappropriate portion of the Stirling cycle. The cooler liner 10702directs the flow of the working gas along the inner surface of thecooler liner 10702. An improvement of the presently described engine isan increased heat transfer surface area in the heat exchanger where theouter diameter of the cooler liner 10702 is reduced to increase theinterfacial area 10706 with a plurality of extended surfaces, forinstance, longitudinal arranged fins 10707, or pins, provided around theoutside diameter of the cooler liner 10702 and extending into theinterfacial area 10706 between the inner surface of the water jacketsleeve 10704 to increase the surface area of the heat exchangingsurfaces and provide more efficient cooling of the working gas/fluid.The outer diameter of the cooler liner wall 10708 can be reduced to anextent so that the cooler liner wall 10708 is relatively thin, ascompared to the radial length of the longitudinal fins 10707, or pins inthe interfacial region 10706 between the cooler liner 10702 and theinner surface of the heat exchanger 10705. The inner wall 10708 of thecooler liner is generally maintained an appropriate diameter toaccommodate the working gas flow from the heater head and cylinder. Theinner diameter of the liner is provided with axially arranged fins 10707to direct the flow of gas along the inner wall of the liner andfacilitate the transfer of heat out of the working gas, through thecooler liner and into the coolant water.

It is also important to ensure that the stationary seals utilized in theheat exchanger are to the extent possible redundant and not compromised,particularly where the water jacket sleeve 10704 and cooler liner 10702should sufficiently maintain the coolant water in the interfacial region10706 between these elements and the working fluid inside the coolantline 10702 r. As shown in FIG. 50B, the heat exchanger 10705 in thepresent embodiment has an outer surface which abuts the inner surface ofthe vessel and is sealed with respect to the vessel by an upperstationary seal 10710 and a lower stationary seal 10711. Similarly thecooler liner 10702 inside the heat exchanger 10705 is sealed withrespect to the inner surface of the heat exchanger by an upper seal10713 and a lower seal 10715. A top surface of each of the cooler liner10702 and the heat exchanger 10705 are formed and both support the baseof the heater heads 10703 and provide the communicating interface forthe working gas between the heater tubes 10709 and the heat exchanger10705. An additional or redundant seal can be added at the intersectionbetween the cooler liner 10702 and the heat exchanger 10705 adjacent thetop surface of each element which supports the heater heads 10703. Thisredundant seal 10712, for example a 45 degree o-ring, is located axiallyspaced above the upper stationary cooler liner sealer 10710 and extendscircumferentially around the entire joint between the heat exchanger10705 and the cooler liner 10702. The addition of the heater head baseas it is supported on the top surfaces of the liner 10702 and heatexchanger 10705 compresses the redundant seal into the joint and addsredundancy to the system to prevent the escape of cooling water and/orworking gas/fluid from the working space.

In a further embodiment, as seen in FIGS. 53C and 53D, the heatexchanger 10720 has a circumferentially disposed interfacial area 10722comprising a plurality of longitudinally extending tubes 10724 providedaround the outside diameter of the heat exchanger body 10726. The tubes10724 extend through the interfacial area 10722 to increase the surfacearea of the heat exchanging surfaces and provide more efficient coolingof the working gas/fluid. This is accomplished by routing the workingfluid through the tubes 10724 that are surrounded by the cooling fluidin the interfacial area 10722 instead of on the opposite side of thecooler liner as in the above described embodiments. The tubes 10724 maybe assembled onto the heat exchanger 10720 through a brazing processwherein the completed assembly is run through a brazing oven to solidifythe connection between the tubes 10724 and the heat exchanger body10726.

In the current embodiment, the tubes 10724 and the heat exchanger body10726 are constructed of the same material to simplify the assemblyprocess. In one example, the tubes 10724 and heat exchanger body 10726are fabricated from 300 series stainless steel. In another example, thetubes 10724 and exchanger body 10726 are fabricated from an aluminumalloy such as one of the following but not limited to AL7075-T6,

In an alternate embodiment, the tubes 10724 are constructed of a 300series stainless steel while the body 10726 is constructed of a 400series stainless steel, such that the tubes 10724 have a lowercoefficient of thermal expansion. The tubes' 10724 lower coefficient ofthermal expansion may cause the tubes to not expand as much during theheating period as the body 10726. The braze will join the tubes 10724 tothe body 10726 at the highest cycle temperature, when the tubes have notlengthened as much as the body. Subsequently during cooling the heatertubes will be compressed as the length of the body is reduced by morethan the tubes. However, as the tubes are soft due to the hightemperature, some or all of the tubes may buckle slightly due to thecompression force. The buckling is not enough to weaken the tube orrestrict flow through the tube. One benefit is that during operation,when the body 10726 and tubes 10724 are heated to a lower temperature,the greater thermal expansion of the body will not break tubing/bodybraze joint because the buckled tubes provide structural elasticity. Inother words, during use, unbuckled tubes 10724 are colder, less elasticand thus can apply a greater and repeated load on the braze joint whichcan lead to failure. The buckled tube is less stiff, due to the newshape of the tube and applies a lower load on the braze joint duringuse, which may lead to less failures and longer periods betweenfailures. and during use at lower temperatures, the now stronger andless elastic tubes. In an embodiment, the heat exchanger assembly may gothrough a second heating phase and possibly a second brazing after theinitial brazing process. This second heating phase or brazing will againjoin the tubes 10724 to the body 10726, when the tubes 10724 have notextended as much as the body 10726. The cooling process from a brazingtemperature which is above the annealing temperature may allowlongitudinally compressed tubes 10724 to slightly deform away from acompletely vertical structure, as in a slight bend, such that thedeformation eliminates any pre-compression in each tube 10724.

In a further improvement to the drive system a more easily constructedand easy to maintain connection between the link rod 10826 and therocking beam 10816 is described. Referring now to FIG. 51A, a rockingbeam drive mechanism 10801 is shown. In this embodiment, the rockingbeam drive mechanism has pistons 10802 and 10804 coupled to two rockingbeam drives 10801. In the exemplary embodiment shown more clearly inFIG. 51B, the link rod 10826 is coupled at a first end to the piston rodvia a link rod upper pin 10832, and a second end of the link rod 10826may be coupled to one end of a link rod lower pin 10832 attached to theyolk of the rocking beam 10816. The link rod lower pin 10832 had beenpreviously accomplished by press fitting a pin 10823 into a passage ofthe link rod 10826, and with bearings provided on either side of thelink rod 10826 and around the pin 10823, the second end of the link rodis secured to the rocking beam drive 10801 in a yolk 10825. The pin10823 extends into respective pin passages in the yolk 10825 of therocking beam 10816 in order to complete the link rod lower pin 10832structure. A bearing is also provided between the pin passages in therocking beam 10816 and the pin 10823 to facilitate the pivoting of thelink rod and pin relative to the rocking beam 10816.

The present embodiment eliminates the need for a press fit of the pininto the passage in the second end of the link rod 10825. The press fitmade it difficult to maintain, fix assemble and disassemble in anymanner this end pivot structure during maintenance of the engine. Asseen in FIG. 51B the link rod lower pin 10823 is provided to be insertedwith a loose fit into and through the passage in the second end of thelink rod 10826. A bearing 10822 may be provided around the link rodlower pin 10823 on either side of the link rod 10826, and the width ofthe bearing 10822 is reduced in order to fit a retaining ring 10828 ontothe pin 10823 adjacent each bearing to retain the pin 10823 and bearingsaxially aligned in the passage of the link rod 10826 and in the yolk10826. With the pin 10823 and bearings essentially axially fixed by theretaining rings 10828 to the link rod 10826, the pin 10823 and link rodpassage can have a loose fit so that the pin 10823 can be easily removedfrom the link rod 10826, when disassembly is necessary, merely byremoving the retaining rings 10828 and sliding the pin 10823 out of thelink rod passage. A lubricating oil passage 10829 may be provided in thelink rod lower pin 10823 to communicate with a oil passage 10838 in thelink rod 10826 and provide oil to the bearings 10822 and the respectivesurface of these pivoting components.

The link rod upper pin 10832 is similarly arranged with a loose fit withthe first or upper end of the link rod 10826. A bearing 10834 in thiscase is provided directly between the bearing surface of the upper pin10832 and an inner surface of the upper link rod passage. A pair ofretaining rings 10836 are applied to grooves in the ends of the upperpin 10832 to maintain the pin in its axial placement in the cross head10840. The bearing 10834 and respective bearing surfaces can be suppliedwith lubricating oil via the oil passage 10838 in the link rod 10826

The arrangement of the Stirling machine discussed above is generallyreferred to and shown as having a vertical orientation, i.e. with thepistons reciprocating generally perpendicularly aligned relative to ahorizontal support surface or ground surface. In another embodiment ofthe present Stirling cycle engine 10903 shown in FIGS. 52A and 52B theengine may be horizontally arranged, i.e. with the pistons 10905, pistonrods 10907, heater heads 10911, cross heads 10913 etc., being arrangedand reciprocating in a horizontal orientation relative to a groundsupport surface as opposed to the vertical orientation discussed above.One of the significant challenges in such a design is the arrangementand structure of the oil cooling system in the crankcase 10915 where itimperative to ensure that the mechanical elements of such a horizontalcrankcase such as the cross heads 10913, rocking beam 10919 and othercrankcase components and drive elements are sufficiently supplied with afree flow of oil through the crankcase and back to the oil sump andpump.

As seen in FIG. 52B and by way of general example, the oil coolingsystem comprises a central oil supply line 10921 disseminating a flow ofoil directly to each of the cross head bores 10923 through radial oilpassages 10925. Oil drains down by gravity in the crankcase 10915 intooil sump 10931 which can then be re-circulated back to the central oilsupply line 10921 via a pump 10935 through main line 10937 whichcommunicates eventually with central oil supply line 10921. It is to beappreciated that other oil supply arrangements and orientations can alsobe accomplished, and that the embodiment described with respect to FIGS.52A-B and the horizontal arrangement of the engine and crankcasecomponents is merely exemplary with respect to these figures.

In another embodiment of the engine it is also beneficial to cool thecrankcase by cooling the oil in the crankcase. An oil cooler 10941 showndiagrammatically in FIG. 52B is designed to pick up a substantial amountof the heat generated in the crankcase, and with a co-axial (or atube-in-tube) heat exchanger 11043 shown specifically in FIGS. 53A-B,oil from the crankcase passes through an outer oil channel 11045 over aseries of fins 10947 positioned along the outer surface of a coolingtube 11049 containing flowing cooling water from a cool water source11046. The fins 11047 can be radial fins or axially aligned finsrelative to the cooling tube 11049 depending upon the necessity for adesired oil flow along the outer surface of the coolant tube. Aftertaking up heat from the oil, the cooled oil returns to the main line11037 and the heated water can be dumped to a heat sink 11051. Methodsor apparatus are disclosed for external starters, both manual andpowered, in U.S. patent application Ser. No. 13/447,990, filed Apr. 16,2012 and entitled Stirling Cycle Machine (Attorney Docket No. 184),which is hereby incorporated herein by reference in its entirety.

B-Burner

FIGS. 54-60 disclose a further embodiment of a burner 11201 for use inconjunction with a multiple heater head and piston engine describedpreviously in FIGS. 4, 35, 36A. The present burner 11201 is specificallydirected to the independent heating of multiple heater heads, in thiscase four (4) heater heads, each heated by an individual burner andflame and having a single air inlet 11223, single outer wall 11212, andtwo exhaust openings 11225.

Turning to FIG. 55 the four burner design 11301 of the presentembodiment includes a single blower B providing air for the fuel/airmixture in the ignition process of all the burner head assemblies 11305as shown in FIG. 55. The heater heads 11303 as also discussed above, maybe any of the various embodiments described in the preceding sections,including, but not limited to, tube heater heads, or pin or fin heaterheads as disclosed in U.S. patent application Ser. No. 13/447,990, filedApr. 16, 2012 and entitled Stirling Cycle Machine (Attorney Docket No.184), which is hereby incorporated herein by reference in its entirety.By way of example, the present embodiment is contemplated utilizingheater tubes 11309 through which flow the working gas, for examplehelium, which must be heated by the burner head assemblies 11305 duringthe appropriate portion of the Stirling cycle.

By way of more detailed description and referring back to FIG. 54 aswell, the burner 11301 includes multiple burner head assemblies 11305,one for each of the heater heads 11303, in the case of the presentembodiment there are four (4) heater heads and hence four (4) burnerhead assemblies 11305. The cross-section of FIG. 55 shows three (3) ofthe burner head assemblies 11305. Generally, the burner 11301 is definedby a burner housing as shown in FIG. 54 having a substantiallycylindrical outer wall 11312, although other geometrical configurationscould be imagined. The blower B pumps air into the burner 11301 throughair intake 11223 for purposes of ignition and combustion, and exhaustgases are ejected from the burner via the two exhaust outlets 11225adjacent the base of the burner.

Turning to FIGS. 56 and 57 a top surface 11413 of the burner housingincludes a number of ports 11415 for receiving fuel inputs 11416,igniters 11427, flame and possibly temperature sensors or flame viewingelements. The ports 11415 also facilitate access to a particular burnerhead 11405, as discussed in detail below, without having to remove theentire burner 11401 from the vessel stack-up for maintenance. As seen inFIG. 57, associated with each burner port 11515 on the top surface 11513of the burner is a secondary port 11517 which can serve a number ofpurposes for instance a flame viewing element such as a viewing windowfor viewing the flame of the burner head, or alternatively a spark plug11520 for igniting the fuel/air mixture and/or a sensor for sensing UVlight used in flame detection.

The base of the burner 11601 best seen in FIGS. 58 and 59, is providedwith heater head openings 11619 to accommodate the entrance of theheater heads and respective heater tubes since the burner as a whole isset over and stacked up on a cooling plate 11604 of the vessel so thatthe heater heads 11603 are within and substantially sealed inside orencompassed by a lower region of the burner 11601. The base of theburner 11701 is secured to the mounting plate 11704 in the vesselstack-up by a circumferential band clamp 11710, such as a Marmon clamp,which is provided for securing and critically circumferentiallycentering the burner relative to the cooling plate and lower stack-up ofthe pressure vessel. The centering of each burner head assemblies 11705relative to each of the associated heater heads 11703 is criticalbecause if the flame from the burner head assembly is nearer one side ofthe heater heads 11703 and heater tubes 11709 than another, there willbe not only inefficient heating of the working gas/fluid in the heatertubes 11709, where one set of heater tubes is heated to a highertemperature than other tubes.

The clamp 11710 extends circumferentially and radially around the entirebase of the burner 11701 and provides both radial and axial compressiveforces between the burner base plate and the mounting plate to ensurethat there is both a critical axial sealing pressure to contain the hotexhaust gases in the burner and a radial circumferential alignment ofthe burner heads with the heater heads. The base of the burner housing11711 may be provided in this regard with a circular sealing edge 11721as shown in FIG. 59 which is angled relative to the vertical axialarrangement of the vessel stack-up to create the axial compressive forceand for mateably engaging with an oppositely angled circular sealingedge 11722 of the mounting plate. The circumferential clamp 11710 andthe mating angled circular sealing edges 11721, 11722 of both the burnerand the cooling plate ensure the critical circumferential, i.e. radialalignment of the burner housing 11711 and burner head assemblies 11705with the mounting plate and heater heads 11703 in the vessel stack-up sothat the burner head assemblies 11705 are appropriately aligned with theheater heads 11703 and there is sufficient axial force between theburner 11701 and cooling plate 11704 to contain the hot exhaust gasgenerated in the burner 11701. The circular sealing edge 11721 mayinclude a graphite seal (not shown) between the burner and cooling plateto ensure that the hot gases which are at around 1000 C, where the flametemperature is around 1200 C, do not leak out between the burner 11701and the mounting plate 11704.

The single blower B as shown in FIG. 59 provides air into the burnerhousing 11711 adjacent the top surface 11713 for the fuel/air mixture tothe burner head assemblies. The air intake 11723 is provided atessentially a normal angle to the circular burner housing 11711 andprovides air inside the burner 11701 for combustion as described indetail below. This arrangement of the air intake 11723 at a normal angleto the cylindrical burner housing 11711, better seen in FIG. 58,facilitates air entering the burner 11701 with a designed pressure dropwhich is important for incoming air to the burner to maintain a desiredair velocity for maximizing heat transfer efficiencies as the air passesthrough the air intake manifoldand into a preheater 11726 where theincoming (cold) air is warmed by the exiting (hot) exhaust gasses. Asingle blower B is placed in communication with the single air intake11723 to provide air to all the burner head assemblies 11705 in theburner 11701. A pair of exhaust outlets 11725 are also connected normalto the substantially cylindrical burner 11701 and spaced approximately180 degrees apart around the base of the burner 11701. Prior to exitingthrough the exhaust outlets 11725 the exiting exhaust from the burner11701 preheats the incoming air in the preheater 11726 described indetail below, then exits the burner 11701 from one of the two exhaustoutlets 11725.

Observing FIG. 59, each burner head assembly 11705 has a fuel injector11724, an igniter 11727 of one kind or another, for instance a sparkplugor glow-plug, a flame detection device 11729 which may also be providedin the secondary port 11717 as shown. Fuel, either liquid fuel orgaseous fuel is fed to the fuel injector 11724 via a fuel line 11731from a fuel source F and is dispersed as a fine mist or vapor by thenozzle 11734 of the fuel injector 11724 into a prechamber 11728 of theburner head assembly 11705. In the prechamber 11728 the dispersed fuelis combined with a desired volumetric flow of air from the preheater11726, preferably preheated to a desired ignition temperature by theexhaust as discussed in detail below, to form a desirable fuel/airmixture for ignition. The fuel/air mixture then is ignited by theigniter 11727 and combusts at least partly inside the prechamber 11728,but more complete combustion may occur after the fuel/air mixture exits,or is pushed, from the prechamber 11728 through the prechamber nozzle11730 of the prechamber 11728 to form a flame which extends from theprechamber 11728 and is directed into a center combustion chamber insidethe heater tube arrangement of each respective heater head 11703.Exhaust from the combustion in the burner 11701 exits the burner via thepreheater 11726 and exhaust manifold 11714 described in detail below.

In the present embodiment of the burner, the single blower B, shown herediagrammatically, may be incorporated to maintain a consistent averageair ratio supplied to the burner 11701 and hence to each of theindividual burner head assemblies 11705. The blower B pumps air at adesired velocity depending on instructions from a controller C forpurposes of ignition, then once ignition has occurred, the desired airflow rate may be regulated by the controller C dependent on datareceived from sensors including but not limited to an oxygen sensor. Amore complete description of the burner control algorithm is providedbelow. The single blower B is also controlled dependent on the data fromindividual burner head assemblies, for example in the case of at leastone burner head assembly being extinguished or not igniting thecontroller may decrease the blower rate to facilitate ignition in theextinguished burner head assembly. The fuel input may be correspondinglycontrolled in the remaining burner head assemblys 11705 to accommodatesuch an air velocity decrease. In any event, the blower B is intended toprovide a consistent flow rate to each of the multiple burner headassemblys 11705 in the burner after passing through the preheater 11726.An important aspect of the present embodiment is the consistent flow andvelocity of cold air developed by the blower B and the efficient heatingof the incoming air through the extraction of waste heat from in thepreheater 11726, to raise the cold air temperature thereby improving theefficiency of combustion processes and the burner unit.

The blower B connects through the air intake 11723 in the outer wall11712 of the burner housing into a cold air channel 11735 of thepreheater 11726. The cold air channel 11735 extends circumferentiallyaround the burner inside the outer wall 11712 of the burner 11701 anddirects the cold air developed by the blower B down around an insulatedintermediate baffle 11739 and up into the preheater. The intermediatebaffle 11739 is insulated to protect the outer wall 11712 of the burner11701, and anyone or thing that comes in contact with the outer wall,from the intense high temperatures inside the burner 11701. Also, theinsulated baffle 11739 ensures that heat captured by the incoming air inthe preheater 11726 is not lost directly to the outer wall 11712 of thehousing 11711.

The preheater 11726 essentially begins where the cold air from theblower B drops down through the cold air channel 11735 and enters into apreheater channel 11741 in which the cold air is preheated in order toraise its mean temperature which increases the efficiency of the burnercombustion. The preheater channel 11741 is defined by the intermediatebaffle 11739 on one side, and on the inner side, an exhaust manifoldwall 11743. The exhaust manifold wall 11743 directly separates theincoming cold air from the exhaust air exiting the burner and providesfor the heat transfer from the exiting exhaust to the incoming cold airin the preheater channel 11741. The heat transfer efficiency through themanifold wall 11743 in the preheater 11726 is critical because thehotter the incoming cold air can be raised by the preheater 11726, theless fuel is necessary to get the gas up to desired ignition andcombustion temperatures. The preheater channel 11741 also extendscircumferentially around the entire burner 11701 which provides for amaximum surface are which in some embodiments may produce better heatexchange with the exhaust flowing out of the burner through an exhaustchannel 11744. Inside the preheater channel 11741 are a series ofradially extending fins 11745 which are directly connected to theexhaust manifold wall 11743 and assist in efficient heat transfer fromthe exiting exhaust air through the manifold wall 11743 into the air inthe preheater channel 11741. Exhaust side fins 11746 may also beconnected to the exhaust manifold wall 11743 extending into the exhaustchannel 11744.

The cold incoming air is preheated to a desired temperature, forexample, but not limited to 600-750 C, in the preheater 11726 whichfacilitates ignition and combustion as the air is directed to the burnerhead assemblys. The amount of preheating which may be accomplished isprimarily based on the efficiency of heat transfer from the exitingexhaust so that as the exhaust temperature is raised during operation ofthe engine, the incoming cold air can be accordingly preheated to ahigher temperature. The preheated air exits the preheater channel 11741and is directed radially into a hot air chamber 11747 which communicateswith each of the multiple burner head assemblies 11705. It is to beappreciated that the preheated air enters the hot air chamber 11747through a substantially 360 degree circumferential opening around theexit of the preheater channel 11741 so that a consistent flow rate ofpreheated air is delivered to each of the burner head assemblies 11705.While additional channels or passageways (not shown) may be provided inthe hot air chamber 11747 to direct the preheated air in the hot airchamber to a specific burner head, the 360 degree output from thepreheater channel of the present embodiment is important since there isonly one blower B developing the air flow into the engine. In previousengines a multitude of blowers delivered a desired air flow to each ofthe burner head assemblies, for instance where there were four (4)burner head assemblies, there were four (4) blowers, one directed toeach burner head. However, having a blower associated with each burnerhead 11705 on a multiple burner head engine is expensive and adds asignificant amount of weight to the engine. In any event, a singlepreheater is much less expensive and less complicated from a controlstandpoint than separate preheaters of each heater head.

The preheated air is directed in the hot air chamber 11747 to theindividual burner head assemblies 11705 and specifically to anintersection with a nozzle 11734 of each fuel injector 11724 in eachburner head 11705 and an igniter 11727. The fuel injectors 11724 may useeither liquid fuel or gaseous fuel but in either case the fuel isejected from the injector into the prechamber 11728 where the fuel mixeswith the preheated air to attain a desired fuel/air ratio or mixture foreither ignition of the burner head 11705, or, combustion where theburner head 11705 is currently supporting a flame. The fuel injectors00024 inject the fuel into the prechamber 11728 directly below the fuelinjector 11724 and the preheated air is combined in the prechamber 11728with the liquid or gaseous fuel. The fuel may be delivered as a mist orvapor, combined with the preheated air and ignited in the prechamber11728 by the igniter 11727. While ignition of the fuel/air mixture mayoccur to some extent in the prechamber 11728, the flame derived from theignition and combustion of fuel/air in the prechamber 11728 needs to bepushed out of the prechamber 11728 to be more efficient and provide therequisite thermal output. It is preferable that the constant combustionflame which heats the heater heads 11703 and heater tubes 11709 bepushed out of the prechamber 11728 and actually extend beyond the endcone 11730 of the prechamber 11728 and into the combustion chamber11750. This is accomplished by providing appropriate adjustment to thefuel/air mixture by the controller and by the prechamber and nozzlegeometry to properly control the shape of the flame. Structural elementsmay be added to the prechamber to improve the shape of the flame in thecombustion chamber as disclosed in U.S. patent application Ser. No.13/447,990, filed Apr. 16, 2012 and entitled Stirling Cycle Machine(Attorney Docket No. 184), which is hereby incorporated herein byreference in its entirety.

Another aspect of the present embodiment shown in FIG. 60 is theprechamber support which extends around the outermost wall of theprechamber 11928 adjacent the exit cone 11930 and includes a heater headrestrictor 11937 restricting the flame exhaust gases from passing overthe heater head 11903. The restriction increases the pressure drop atthe top of the heater head 11903 and improves heat transfer from the hotgas to the helium inside the heater tubes 11909 by encouraging the hotgas to flow through the heater tubes.

A substantial amount of the heat not used to heat the working fluidremains in the exhaust gases and thus the efficiency of the entireengine can be increased by using the excess exhaust gas heat to preheatthe incoming air. After heating the heater head 11903 and heater tubes11909 the hot combustion gases are forced out an exhaust inlet 11953 bynewer combusted gases into the exhaust channel 11944 defined partly byan inner wall 11942 of the exhaust manifold. The exhaust passes downalong the exhaust channel 11944 exchanging a substantial amount of heatthrough the preheater wall 11943 to the incoming cold air entering thepreheater 11926 via the preheater channel 11941. The exhaust gasesshould flow as quickly as possible through the preheater 11926 as theheat transfer from the exhaust gases is dependent upon the velocity ofthe exhaust. Another aspect of the present embodiment limits thepressure drop of the exhaust gases by allowing the exhaust gases flowout from two exhaust outlets 11925, as opposed to one exhaust outlet,arranged around the bottom of the exhaust manifold 11914. The shorterflow path provided by the two exhaust outlets 11925 for the exhaustleaving the exhaust manifold 11914 lowers the pressure drop for theexhaust, the blower does not have to work as hard, and thus the blowerload is reduced on the engine.

FIG. 60 also discloses the use of spark plugs 11960 in the secondaryport 11917 (rather than a UV viewing window) through the burner housing.In certain cases a flame sensor may also be inserted through thesecondary port 11917 which extends into the burner adjacent each of theprechamber nozzle so that flame detection can occur. In any event, thesecondary port 11917 provides for access to the burner head 11905 sothat a sensor or window, or ignition components for instance glo-plugsor spark plugs as shown here can be inserted down into the burner head11905 to ignite different gases and fuels. Gaseous fuel use maynecessitate the spark plug 11960 to ignite the fuel air mixture adjacentthe nozzle of the prechamber whereas liquid fuel uses glow plugs and aregenerally located closer to the fuel injector itself. In the embodimentshown here a high voltage conductive element 11961 is encased within andinsulative layer 11963 and a ground layer 11965 and inserted through thesecondary port so that the exposed conductive element 11961 is exposedin the combustion chamber to ignite the fuel/air mixture exiting theprechamber 11928.

The ability to see and/or detect each flame is important so that each ofthe four individual burner head assemblies 11905 and respective flamecan be appropriately adjusted by the controller. It is to be appreciatedthat such flame detection and viewing may be accomplished by manyembodiments, including but not limited to an actual viewing window forexample having appropriate lenses in the tube which allow a humanoperator to look through the tube and visually identify a flame withinthe range of visible wavelengths in the combustion chamber.Alternatively, the viewing window may include a camera or other imagedata receiving and recording device such as a UV light sensor anddisplay for visually displaying a received representation of the flamein the combustion chamber. Other types of heat sensors including but notlimited to thermocouples, infrared thermometers, and thermisters, may beused to identify and quantify the flame and flame characteristics in thecombustion chamber.

With only a single blower providing air to four burner head assemblies11905, generally a variable in addition to air, such as fuel, must bealtered to obtain a desired flame quality. Keeping one blower providingair to all four burner heads is especially helpful for cost and forblower power consumption.

With liquid, diesel or other gaseous fuel, the UV viewing window will becompromised because the fuel vapor tends to absorb the UV radiation fromthe flame. Without the UV window as in the previous embodiment it maystill be important to detect the flame and the temperatures in thecombustion chamber. The electrode of the sparkplug may be utilized as asensor in some cases to detect the flame. Such data can be forwarded tothe controller to determine the flame and combustion disposition in thecombustion chamber 11950. Another method of flame detection obtainstemperatures with temperature sensors inside the heater head, forexample a thermocouple attached on the walls of the heater tubes canprovide data to the controller to alter the operational conditions ofthe engine. This temperature data is used to judge the temperatureand/or flame quality based on temperature/flame data and helps thecontroller decide what operational mode, as discussed in further detailbelow, to set for each burner head 11905 and for the engine as a whole.

Burner Control

The burner may be operated in several modes as shown in FIG. 61, andreferring in part to FIG. 60, according to predetermined electronics andsoftware programs embodied in an electronic controller. The operationmodes evaluated by the controller include at least a start-up mode12002, normal operation mode 12004, shut-down mode 12006 and a stop mode12008. The start-up mode includes the initial ignition of a richer fuelmixture in the prechamber 11928 to ease ignition as colder mixtures havea narrower ignition range of fuel/air ratio that are ignitable comparedto the range of fuel/air ratios that maintain combustion. With a desiredfuel/air ignition mixture present in the prechamber 11928, the igniter11927 is actuated and the ignition mixture is ignited. A thermocouple(not shown) in the prechamber 11928 detects what is referred to as adiffusion flame in the prechamber 11928 and once the incoming air is hotenough from the preheater 11926, the flame is pushed out of theprechamber 11928 by either increasing the air flow from the blower B, orincreasing fuel so the flame travels out of the prechamber 11928 andforms in the combustion area adjacent the heater head 11903.

Generally in the start-up mode 12002 as shown in FIG. 61 a user sets adesired blower speed 12003 and fuel/air ratio 12005 for a certain timeperiod 12007, for example 30 seconds. After the predetermined timeperiod the blower shuts off and resets 12011 the start-up phase whichmay include blowing out 12013 any remaining fuel in the engine andexhaust system so that there are no backfires or other damaging eventsfrom residual fuel. The start-up phase may also include for instance anumber of ignition attempts 12009 before resetting and providing theuser wih an error A sensor (not shown) within the prechamber 11928 or avisual sensor using the secondary port 11917 detects if a flame 12010 ispresent within the prechamber 11928 or the combustion chamber 11950. Ifa flame is not detected the system is reset 12011 or if a flame isdetected the temperature readings are taken 12015 from the heater headand oxygen levels are measured 12017 from the exhaust gases. Thefuel/air ratio is then adjusted 12019 based on these readings.

Once the flame is supportable out of the prechamber 11928 and is heatingthe heater head 11903, the control system and operation mode 12004include a number of failsafe triggers 12023 based on sensor data andcontroller evaluation algorithms which evaluate the system and determineif the system should be turned to the shut-down or stop mode. Theoperation mode 12004 monitors levels of heat, power and oxygen forexample and perform shut-down or stopping of the engine, or othermodifications to the system and engine if a temperature reading is toohigh, or exhaust oxygen level is too high or if engine speed exceeds adesired value, or the differential pressure within the air lock is toolow. These are just exemplary triggers for starting shut-down or stopprocedures, other triggers could be used as well or in combination withthese examples.

During normal engine operation, the blower is operated at leastpartially by a control loop which measures the excess oxygen 12017 inthe exhaust to determine blower speed. The failsafe triggers 12023 shownin the flowchart and operation analysis table 12021 in FIG. 61 include:Engine speed exceeds predetermined range; Oxygen levels in exhaustexceed a predetermined range; Generator temperature exceeds apredetermined range; Burner temperature exceeds a predetermined range;Cooler temperature exceeds a predetermined range; Flame/Ignitionfailure; repeatable Failure of flame ignition. It is to be appreciatedthat the described control method is not limited to the disclosedtriggers 12023 and that other triggers, factors and variables may alsobe analyzed by the controller under the start-up and operation modes12002 and 12004.

A failure of the engine in one of these failsafe triggers 12023 directsthe controller C to adjust the fuel/air ratio 12019 and continueacquisition of sensor readings. A preset number of a repeated failure12025 of the engine to run within a predetermined range for any of thesetriggers leads to a shutdown sequence with an immediate fuel turn off12029. The engine however can continue to run in the shut-down mode12006 in many cases. On the other hand, certain events may causecomplete engine stoppage (i.e. shut-off as opposed to shut down) so thatdamage to the engine is minimized. A status check 12037 on systemcomponents is repeatedly run. These shut-off triggers 12034 are forexample, low oil pressure, low airlock pressure differential, and lowengine power levels will ensure complete engine stoppage to preventdamage. During a shut-down mode 12006, the fuel and burner is turned offbut the engine keeps running until the heater head 11903 is cooled to adesired temperature. A system shut-down may also be caused by excessiveheat measurements in a number of components such as the Generator, theburner, or a cooler, or a system shut down may occur if there is afailure to ignite. A shut down due to system failure may trigger a safemode where fuel is pumped out of the system. Any fault or system failureor trigger, will kill the fuel delivery immediately 00036, but theengine will continue to run to cool down the system. The engine runsuntil it reaches a predetermined power level 12035 in the shut down mode12006, or in the event of the more dangerous fail safe triggers theengine is stopped 12008, i.e. the RPMs are set to 0. The shut-down modehelps engine efficiency since the engine, burner and heater heads remainhot for a while, even while there is no fuel supplied, the engine willcontinue to run producing power until the predetermined low power levelis reached. This recovers some of the energy put in at start-up modewhich improves efficiency. Methods to concentrate exhaust gases nearheater tubes is disclosed in U.S. patent application Ser. No.13/447,990, filed Apr. 16, 2012 and entitled Stirling Cycle Machine(Attorney Docket No. 184), which is hereby incorporated herein byreference in its entirety.

Venturi Burner

FIGS. 62A-62D disclose a further embodiment of a burner 12901 for use inconjunction with a multiple heater head and piston engine as describedpreviously in FIGS. 54-60. This embodiment of the burner 12901 is alsospecifically directed to the independent heating of multiple heaterheads 12903. In this embodiment there are four (4) heater heads 12903and respective burner assemblies 12907, although only two (2) arevisible in the cross-section of FIG. 62A, and more or less heater headsare of course also possible for the engine. As in the previouslydiscussed embodiments the heater heads 12903 and burner assemblies 12907are encompassed by a burner housing 12911 and each heater head 12903 isheated by an individual burner assembly 12907 and flame supplied with afuel/air mixture for combustion via a blower B and air inlet 12923 and afuel injector 12927. As described in further detail below the exhaustingcombustion gases are used to pre-heat the incoming air and, followingcombustion, one or more exhaust outlets are provided through exhaustopening(s) 12925 (not shown) to finally exhaust the combustion gasesfrom the burner housing 12911.

More specifically observing FIG. 62A, the four burner design 12901 ofthe present embodiment includes the single blower B connected to thehousing 12911 providing air for the fuel/air mixture in the ignitionprocess of each the burner assemblies 12907 as shown. The heater heads12903 themselves may be any of the various embodiments of tube heaterheads described in the preceding sections, including, but not limitedto, straight tube heater heads or helical tube heater heads as disclosedin U.S. patent application Ser. No. 13/447,990, filed Apr. 16, 2012, andincorporated herein. By way of example, the present embodiment iscontemplated utilizing heater tubes 12909 through which flow the workinggas, for example helium, which is heated by the respective burnerassemblies 12907.

By way of more detailed description, the burner 12901 includes multipleburner assemblies 12907, one for each of the heater heads 12903. In thecase of the present embodiment there are four (4) heater heads 12903 andhence four (4) burner assemblies 12907. FIG. 62C discloses the burnerhousing 12911 having a substantially cylindrical outer wall 12912,although other geometrical configurations could be accomplished. Theblower B pumps air into the burner 12901 through air intake 12923 forpurposes of ignition and combustion, and exhaust gases are ejected fromthe burner via the exhaust outlet 12925 adjacent the base of the burner.

As shown in FIG. 62C, the top of the burner housing provides one or moreports 12915 in the top of the burner 1212A for receiving fuel injectorassemblies 12916. The injector assembly 12916 may provide a port forfuel 12916A, a port to mount an ignitor 12916B and a sensor port 12916Cto monitor the hot combustion air. The ignitor 12918 may be a spark plugand may function as the high voltage electrode of a flame ionizationflame detector. In other embodiments, the ignitor 12918 may be a hotsurface ignitor. In other embodiments the port 12916B may be used for avision based flame detection circuit including but not limited to one ormore of the following: IR flame detector, visible light flame detectorand/or UV flame detector. Injector assembly 12916 may be removablyconnected to burner 12901. This arrangement may be beneficial for manyreasons, including, but not limited to, to allow the injector assembly12916 to be changed, cleaned and/or modified. Fuels with differentenergy densities may require a different sized port 12945.Alternatively, the fuel ports 12945 may be made very small or fittedwith nozzles for liquid fuels. In some embodiments, liquid fuels may beatomized by forcing the liquid through very small orifices. In someembodiments, the liquid fuel may be atomized by a number of nozzlegeometries including, but not limited to, one or more of pressurenozzles, and/or air blast nozzles. In some embodiments, the injectorassembly may be sealed to the burner housing 12911 with exhaust gaskets12922. In some embodiments, the fuel injector assembly 12916 may bebrazed or welded to the burner top 12912A.

In some embodiments, the burner as a whole is set over and stacked up ona cooling plate 12904 of the vessel so that the heater heads 12903 arewithin and/or are substantially sealed inside and/or are encompassed bya lower region of the burner 12901. In some embodiments, the burner base12902 may be secured to the cooling plate mount 12904A in the vesselstack-up by a circumferential band clamp 12910, such as, in someembodiments, a Marmon clamp, which is provided for securing andcritically, circumferentially centering the burner relative to thecooling plate 12904. However, in various embodiments, the burner base12902 may be secured using another securing apparatus/device. In someembodiments, the burner base 12902 may include one or more pins (notshown) that mate to holes in the cooling plate mount 12904A to orientthe burner such that each burner assembly 12907 may be substantiallycentered over each heater head 12903. As previously discussed, thecentering of each burner assembly 12907 relative to each of theassociated heater heads 12903 is critical for many reasons, including,but not limited to, if the flame from the burner assembly is nearer oneside of the heater heads 12903 and heater tubes 12909 than another,there may be inefficient heating of the working gas/fluid in the heatertubes 12909, and a potential for certain heater tubes to be heated to ahigher temperature than other tubes. As discussed in further detailbelow, the efficiency of the engine may be improved when even andconsistent heating of the heater tubes and the working gas/fluid isaccomplished in the engine.

Referring again to FIG. 62A, in various embodiments, each burnerassembly 12907 has a fuel injector assembly 12916, an igniter, which, insome embodiments, may be a sparkplug 12918 or glow-plug or anotherigniter, and a flame detection device which may also be provided in asecondary port. Fuel, which may be, in various embodiments, eitherliquid fuel or gaseous fuel, is fed to the fuel port 1216A via a fuelline from a fuel source F and is dispersed as a fine mist/mist or vaporthrough the multiple fuel ports into an ejector 12941 of the burnerassembly 12907. The ejector 12941 in this embodiment is a venturi typeejector as disclosed for example in U.S. patent application Ser. No.12/829,320 filed Jul. 1, 2010, now U.S. Publication No.US-2011-0011078-A1 published Jan. 20, 2011 and entitled Stirling CycleMachine (Attorney Docket No. 178), which is hereby incorporated hereinby reference in its entirety. In some embodiments the venturi 12941 maybe beneficial for many reasons, including, but not limited to, providingthe benefit of reducing or eliminating the need for a completelyseparate fuel control scheme as regulation of the airflow changes thevacuum which, in turn, correspondingly affects the fuel flow andregulates the burner power. In some embodiments, the venturi allows useof typical gas pressures in buildings, e.g., 7 inches of water column,without a compressor. The blower B forces air into an initial swirlerportion 12943 of the venturi 12941. The flow of swirled air through theventuri draws in a proportional amount of fuel through the fuel inletports 12945 on the fuel injector assembly 12916 which, in variousembodiments, may be positioned at, or in, the venturi throat 12947. Thisswirling fuel/air mixture exits the venturi and forms a swirl stabilizedflame in the combustion chamber 12931.

The fuel/air mixture is ignited by the igniter or spark plug 12918 andmay combust partly inside the diverging section of the venturi 12949.More complete combustion may occur in the combustion chamber 12931extending down inside the heater tube 12909 arrangement of eachrespective heater head 12903. The hot combustion gas then passes betweenthe heater tubes 12909 and collects on the outside of the heater heads12903. In various embodiments, exhaust shields 12908 direct the hotcombustion gases to flow through the outer row of heater tubes 12909.The now cooled exhaust exits the burner via the preheater and exhaustmanifold described in greater detail below.

In various embodiments of the burner 12901, the single blower B, shownhere diagrammatically, may be incorporated to maintain a consistentaverage air flow supplied to the burner 12901 and hence to each of theindividual burner assemblies 12907. In some embodiments, the blower Bmay also produce a variable air flow when necessary to control thefuel/air mixture in the venturi 12941. The blower B may pump air at adesired velocity depending on instructions from a controller forpurposes of ignition, then once ignition has occurred, the desired airflow rate may be regulated by the controller dependent on data receivedfrom sensors, which may include, but not limited to, an oxygen sensorsampling the cooled exhaust gas. In some embodiments, a fuel system withvariable flow valves for each heater head, which, in some embodiments,may be a Maxitrol EXA-40 valves for example, as manufactured by theMaxitrol Company, Southfield, Mich., USA, may control the fuel flow toachieve the commanded temperature on each heater head. The blower may beadjusted to achieve the desired fuel/air ratio.

An important aspect of the present embodiment is the efficient heatingof the incoming air through the extraction of waste heat in the exhaustto raise the incoming air temperature thereby improving the efficiencyof combustion processes and the burner 12901. The blower B connectsthrough the air inlet 12923 in the outer wall 12912 of the burnerhousing into an air channel 12951. The air channel 12951 extendscircumferentially around the burner inside the outer wall 12912 of theburner 12901 and directs the air developed by the blower B across anintermediate baffle 12953 and up into the hot manifold 12957 beforeentering the swirlers 12943 of the venturi ejectors 12941. Theintermediate baffle 12953 directly separates the incoming air from theexhaust gases exiting the burner through an exhaust channel 12955 andprovides for the heat transfer from the exiting exhaust to the incomingair in the air channel 12951. The heat transfer efficiency across theintermediate baffle 12953 is critical because the hotter the incomingair can be heated, the less fuel is necessary to reach the desiredcombustion temperatures.

The incoming air is preheated to a desired temperature, for example, butnot limited to 600-750° C. These are many reasons preheating theincoming air may be beneficial and these include, but are not limitedto, facilitating ignition and combustion as the air is directed to theburner assemblies 12907 and/or increasing the thermal efficiency of theburner by capturing some of the thermal energy in the combustion gasesexiting the heater heads. In some embodiments, preheating of the air mayreduce the hot exhaust temperature from 900° C. to 300° C. In someembodiments, the amount of preheating which may be accomplished may berelated to the efficiency of heat transfer from the exiting exhaust tothe incoming air. The heat transfer across the intermediate baffle 12953may be improved by adding rows of folded fins 12952 on the air side andfolded fins 12952 a on the exhaust side of the intermediate baffle12953. The folded fins may be brazed to the intermediate baffle 12953 toassure good thermal attachment. In various embodiments, the materialproperties of the folded fins may be optimized for the operatingtemperature. For example, in some embodiments, the rows of folded finsnear the top may be heat resistant metals, which may include, but is notlimited to, INCONEL 625, while lower and cooler folded fins may havehigher thermal conductivity but lower operating temperature. In variousembodiments, the materials for these folded fins may include, but is notlimited to, stainless steel 409 or Ni 201 for example. The preheated airexits the air channel 12951 and is directed radially into a hot airmanifold 12957 which communicates with each of the multiple burnerassemblies 12907 specifically directing the preheated air to the swirlerportion 12943 of the venturi 12941. In various embodiments, thepreheated air enters the hot air chamber 12957 through a substantially360 degree circumferential opening around the exit of the air channel12951. In some embodiments, this may result in a consistent flow rate ofpreheated air delivered to each of the burner assemblies 12907. Invarious embodiments, additional channels or passageways (not shown) maybe provided in the hot air chamber 12957 to direct the preheated air inthe hot air chamber to a specific burner head. In various embodiments,the 360 degree output from the air channel 12951 is used when there isonly one blower B developing the air flow into the engine.

Airlock and Working Fluid Repressurization System

As described previously in this application, the power, life and valueof a Stirling engine may be maximized, in some embodiments, by buildingan oil lubricated drive contained in a pressure vessel, generallyreferred to herein as the crankcase, and sealing the working space ofthe Stirling engine which contains the working fluid, for examplehelium, from the crankcase oil with flexible membranes or bellows suchas the rolling diaphragms also discussed above. The rolling diaphragmsattach to the moving piston rod and the engine casing structure enablethe piston rod to move relative to the casing and to provide an oiltight seal between the oil filled crankcase and the workspace, ensuringthat the lubricant is maintained in the crankcase and does not disperseinto the working fluid of the Stirling. Dispersion of the oil from thecrankcase into the working fluid would lead to engine failure. In orderfor the bellows to function for thousands and millions of cycles, asnecessary, a small pressure difference must be maintained across thebellows. An airlock is provided between the constant pressure crankcaseand ossilating pressure workspace to create a volume at the meanpressure of the workspace. The pressure of this airlock may becontrolled to provide a constant pressure difference across the bellows.This is described in more detail above.

An important aspect of the rolling diaphragm and oil lubricatedcrankcase relates to the use of an airlock 10401 and an airlock pressureregulation system 10411 as shown previously in FIGS. 44A and 44B. By wayof review, the airlock pressure regulation system 10411 provides thebenefit of ensuring that an appropriate/desired pressure differential ismaintained across the rolling diaphragms 10490 and that working gasescaping into the crankcase is cleaned of lubricating oil returned tothe working space.

Referring now to FIG. 63, as previously discussed, the pressure ofairlock space 13101 is desired to be maintained at essentially 1500 PSIand equal to the mean pressure of working space 13103. Other pressuresare of course possible with 1500 psi being an example of one embodimentof the airlock. In various embodiments, the pressure in the workingspace 13103 may vary approximately +/−300 psi so the function of theairlock space 13101 is to insulate the diaphragms 13190 from suchfluctuations and maintain itself at around the necessary pressure, byway of example here 1500 psi, relative to the 1485 PSI charged in thecrankcase 13110 so that there is approximately about a 15 psi differencebetween airlock space 13101 and crankcase space 13110. In variousembodiments, the crankcase, workspace and air lock are initiallycharged, at room temperature, to a pressure well below the desiredoperating pressures because the pressure rises as the gas in the threevolumes, i.e., workspace, crankcase and air lock, heat up during enginestartup. In addition, during different operating conditions, thetemperature of one or more of the three volumes may change causingchanges in pressure across the bellows. In some embodiments, an Airlockdelta Pressure Regulation (AdPR) block 13111 is provided between thecrankcase 13110 and the airlock space 13101 to create and maintain theexemplary 15 psi pressure differential (and/or the desired pressuredifferential) therebetween in all operating conditions including but notlimited to engine starup, engine shutdown, changes in temperature orspeed, leaks from one volume to another or leaks from one volume toambient.

The embodiment as shown in FIGS. 63 and 132 A-C is referred tohereinafter in general as the “AdPR block” or “AdPR system”. In theseembodiments the AdPR block 13111 is connected between the crankcase13110 and the airlock 13101. The AdPR block regulates the pressuredifference between the airlock 13101 and the crankcase 13110. When thereciprocating pistons 13124 of the Stirling cycle machine are moving,the AdPR block 13111 keeps the airlock pressure preferably 10 to 15 PSIabove the crankcase pressure so that the rolling diaphragms aremaintained in a desired arrangement, essentially bellowing into thecrankcase. It is to be appreciated that a range of 5 to 20 PSI ispossible and other pressure differentials can be accomplished by theregulator as well. The desired pressure difference across the bellowseal may depend on the material and physical dimensions of the bellows,so that other bellows may require other pressure differences. While theStirling cycle machine engine is off, in various embodiments, the AdPRblock 13111 keeps the airlock pressure preferably less than 15 PSI abovethe crankcase pressure and not more than 5 PSI below crankcase pressure.It is permissible to have a greater fluctuation of pressure differentialwhen the engine is off since there are little or no dynamic forces beingapplied to the rolling diaphragms 13190 via moving piston rods. Invarious embodiments, the desired pressure difference may vary.

In some embodiments of the AdPR block 13111, for example, as shown inFIGS. 64A-C, the AdPR block 13111 has an outer housing 13207 havingmounting brackets 13229 or other attachment fixtures to secure the AdPRblock 13211 on or near to the Stirling cycle machine. Along the centerof the housing 13207 one or more ports may be positioned with at leastone port connected to the crankcase designated as crankcase port 13251and at least one port connected to the airlock designated as airlockport 13249. The crankcase port 13251 is connected to the oil filtervolume 13219. The airlock port 13249 is connected to the AdPR airlockspace 13202 surrounding the pump 13221 (FIGS. 64D, 64E, 64F).

A port 13457 in FIGS. 65A, 65B connects the AdPR airlock space with theairlock side of the spool valve 13342. A working gas fill port 13214 anddrain port 13215 may be positioned as shown at either end of the block13211 with a first end of the housing having an electrical conduit andwire feed-thru 13317 necessary to drive a pump motor 13313 shown inFIGS. 65C, 65D.

Turning to the cross-section seen in FIG. 65B, in various embodiments,the AdPR block 13311 may include a spool valve regulator 13341 withappropriate passages 13353, 13344 to the pump (not shown) and componentsof a linear position sensor 13352. The passage labeled 13344 operates atapproximately the crankcase pressure and connects the spool valve to thepump inlet port 13439 as shown in FIGS. 65G, 65H. The passage labeled13353 operates at approximately the airlock pressure and connects thespool valve to the pump outlet port 13438 as shown in FIGS. 65G, 65J. AnAdPR pump controller 13350 is provided in some embodiments and in someembodiments, the AdPR pump controller 13350 may prevent premature wearof the pump components and reduce airlock pressure variations. The AdPRpump controller 13350 communicates with at least the spool positionsensor elements 13352 and may operate the pump only when needed toincrease the pressure difference between the crankcase volume 13110 andthe airlock volume 13101 as sensed by the position of the spool valve.In some embodiments, the AdPR controller 13350 may stop the pump 13312when the spool position indicates that the air lock pressure issufficiently above the crankcase pressure. In one example, the AdPRcontroller 13350 may change the speed of the pump 13321 proportionallyto the spool valve position in order to achieve a more constant pressuredifference across the bellows 13190. In an example, the AdPR controller13350 may run the pump at maximum speed for spool valve positions beyonda given value. In an example the AdPR may command the engine to zerospeed for spool positions beyond a second given value where that levelindicates the airlock pressure is not sufficiently greater than thecrankcase pressure. In various embodiments, the position sensor 13352may be a proximity sensor that senses an absolute position or, may be arelative, i.e. differential, position sensor. In various embodiments,the sensor may be any other type of sensors. Depending on the locationof the spool 13342, the position sensor 13352 may sense the position ofthe spool 13342 and transmits that data to the controller 13350.

In some embodiments, the position sensor 13352, shown herein, may be anLVDT (Linear Variable Differential Transducer) linear position sensorwhich is a type of electrical transformer used for measuring lineardisplacement. Various embodiments of an LVDT generally have threesolenoidal coils (not shown) placed end-to-end around a tube. The centercoil is the primary, and the two outer coils are the secondaries. Acylindrical ferromagnetic core 13354, attached to the object whoseposition is to be measured, slides along the axis of the tube 13356. Analternating current is driven through the primary, causing a voltage tobe induced in each secondary proportional to its mutual inductance withthe primary. The frequency is usually in the range of about 1 to 10 kHz.As the core 13354 moves, these mutual inductances change, causing thevoltages induced in the secondaries to change. The coils are connectedin reverse series, so that the output voltage is the difference (hence“differential”) between the two secondary voltages.

By way of further explanation, in some embodiments, when the core 13354is in its central position, e.g., equidistant between the twosecondaries, equal but opposite voltages are induced in these two coils,so the output voltage is zero. If the core 13354 is displaced in onedirection, the voltage in one coil increases as the other decreases,causing the output voltage to increase from zero to a maximum. Thisvoltage is in phase with the primary voltage. When the core 13354 movesin the other direction, the output voltage also increases from zero to amaximum, but its phase is opposite to that of the primary. The magnitudeof the output voltage is proportional to the distance moved by the core(up to its limit of travel), which is why the device may be described as“linear”. The phase of the voltage indicates the direction of thedisplacement. Because the sliding core 13354 does not touch the insideof the tube 13356, it may move without friction, making the LVDT 13352 ahighly reliable device. The absence of any sliding or rotating contactsallows the LVDT 13352 to be completely sealed against the environment.

Turning to cross-section shown in FIG. 65C, in various embodiments, anelectric motor 13313 may be used to drive the diaphragm pump 13321 that,depending on the position of the spool 13342, may either port to theairlock 13101 or be deadheading. In some embodiments, the seals on thespool are removed so that the pump always pumps to the air lock 13101.In these alternative embodiments, the AdPR controller varies the speedof the pump to produce the same effect on the airlock pressure asdeadheading the pump. The spool valve 13341 comprises again similarcomponents to those described previously in relation to FIGS. 45 and46A-E. Here the spool 13342 is balanced against the pressure of theairlock by a spring 13343. A porting manifold 13302 is shown here toport the pump 13321 from the crankcase side of the spool valve 13342,and into the airlock side of the spool valve 13342. Additionally, theporting manifold 13302 mounts the pump head 13321 and LVDT positionsensor 13352. The crankshaft 13322 and connecting rod assembly 13323connects the electric motor 13313 to drive the pump 13321 and an oilscrubbing filter 13318 is provided adjacent the crankcase port 13351 toensure that any oil which makes its way into the AdPR block 13311 viathe crankcase port 13351 is filtered from the working fluid being pumpedthrough the regulator 13341 to the airlock 13101.

FIGS. 65E-F disclose a still further cross-section of an embodimentdetailing the drain port 13415 which connects the volumes of the AdPRblock 13411 and also connects the engine crankcase 13110 through the oilscrubbing filter 13418 to release pressure in the crankcase 13110,workspace 13103, airlock 13101 and AdPR block 13411. Gas from thecrankcase, which may carry oil and other particles, enters the AdPR viaport 13451 and flows through the oil filter 13418 before entering anyother volumes of the AdPR block 13411. The crankcase gas is filtered toprevent oil and contaminants from entering the pump 1342 and airlock13101 and workspace 13103. The filtered gas at the crankcase pressureflows into the crankcase side of the spool valve 13442 via line 13446.Filtering the gas from the crankcase allows the use of gas from thecrankcase 13110 to maintain the airlock pressure above the crankcasepressure. In some embodiments, an internal valve may be positionedwithin the drain line 13446 to directly connect the airlock 13401 andthe crankcase 13110 through the oil filter 13418 to reduce differentialpressure fluctuations during fill and drain operations and when thepiston rods 13124 are not moving.

Although in some embodiments the AdPR may regulate differentialpressure, the rolling diaphragms 13190 may, in some embodiments, stillexperience fluctuations, which may be especially large during normalfill or drain operation. In some embodiments, alternatively to aninternal valve, a crossover valve (not shown) may be located externallyto the AdPR block. The valve may connect the two sides of the AdPRbetween the airlock 13401 and the crankcase 13110 to avoid a largepressure differential during a fill or drain cycle of the Stirlingengine. The valve may be opened during a fill or drain cycle to greatlyreduce the magnitude of these fluctuations and be closed during normaloperation of the Stirling engine.

In the cross-section in FIG. 66A, the pump port 13555, connecting thediaphragm pump directly to the airlock 13101 via port 13449, may becontrolled by the spool valve regulator 13541. A port 13553 connects tothe pump outlet port 13438 via a check valve 13445 that redundantlyprevents a leak from the airlock 13101 to the crankcase pressure side ofthe AdPR 13411. A port 13544 connecting the crankcase side of the spoolvalve to the diaphragm pump inlet 13448 is also shown.

The spool valve regulator 13541 operation is now described withreference to FIGS. 66A-E. In various embodiments, the spool valve 13542is biased against the airlock pressure so that where the pressuredifference between the crankcase and the airlock is within normallimits, e.g. 15 PSI, as shown in FIG. 63; the airlock port 13553 isclosed by the spool 13542. Additionally, the position sensor 13552 tellsthe pump 13421 that no pumping operation is necessary and the pump doesnot operate. If the airlock pressure drops too low as shown by the valvespool position in FIG. 66A, the spool 13542 is biased by the spring13543 to the left, which permits communication between the airlock port13553 and the pump port 13553. The position sensor 13552 tells the pump13421 to turn on and pump working fluid from the crankcase 13110 to theairlock 13101. The fluid travels first through port 13451, then throughpassages 13446 and 13444, through the pump 13421, past the check valve13445, into the AdPR airlock 13402 via the passage 13453, and finallyinto the engine airlock 13101 through port 13449

In some embodiments, in order to get back into the Stirling cycle enginehelium flows from the crankcase 13110 through the filter 13518 and thesmall amounts of oil in the helium are scrubbed out to keep any oil fromgetting into the AdPR block 13511, the pump 13421 and the Stirlingengine where oil can damage the Stirling engine. In some embodiments,the Stirling engine itself may be disabled to ensure that until thepressure between the crankcase and the airlock is better equalized sothat the rolling diaphragms will not be stressed. In some embodiments,the engine may be stopped if the pressure difference between the airlockand the crankcase is too low as measured by the position of the spoolposition sensor 13552.

In some embodiments, where the airlock pressure is too high as in FIG.66D, the airlock port 13553 may be connected to the crankcase port 13544and the pump 13421 may be disabled while the airlock pressure isreduced. FIGS. 66B and 66C show the airlock pressure within normallimits. In FIG. 66B, the airlock port 13553 is closed by the spool 13541and the spool is still displaced enough according to the LVDT sensor13552 to cause operation of the pump 13421 even without flow from thepump to the airlock. In some embodiments, within normal limits, FIG. 66Cshows the airlock port 13553 closed by the spool 13541 and the spooldisplaced so that the LVDT sensor 13552 does not turn on the pump 13421.

In some embodiments, as shown in FIG. 66E, where the Stirling cycleengine is shut down or in the case of an airlock leak, the crankcase13110 may be pressurized higher than airlock 13101 and workspace 13103,forcing the rolling diaphragms in a way opposite from their intendeduse. In some embodiments, in the event of such a leak, where an internalpressure measurement within the spool is 0-5 PSI higher than the airlockpressure, an internal spool valve seated within the main spool valve13541 may open to equalize the pressure. As with the outer spool valve,in some embodiments, a spring balances against the pressure differential(against the crankcase pressure in this case) and would only open duringa time when the engine is pressurized and off in order to reduce damageto the rolling diaphragms.

In some embodiments, the control of small pressure changes within theairlock and maintaining pressure differential of 5 to 20 PSI above thecrankcase pressure may be achieved using a pump controller that mayaccurately vary the speed of the pump to run faster-slower whennecessary. In some embodiments, using the LVDT sensor and a suitablepump controller, a desired range and/or threshold range may bedetermined. In some embodiments, where the pressure differential isoutside the desired/threshold range, the pump may cycle at a higherrate. In some embodiments, where the differential is closer to or withinthe desired/threshold range, the pump may cycle at a slower rate therebyaccurately controlling pressure within the airlock.

In some embodiments, one or more of the functional components of theADPR module may be located within the Stirling Engine pressure vessel.In one example, the oil filter may be located between the four pistonrods and either in the airlock or just below the airlock in thecrankcase. The inlet line to the pump would run from the clean side ofthe oil filter to the pump location either in an external ADPR or to anelectric pump located within the pressure vessel.

Stirling Engine Controller

Another important aspect of the present embodiment from the standpointof controlling the actuation of the above described airlock and AdPRblock, as well as the rest of the Stirling engine, is the Stirlingengine controller 13660 shown diagrammatically in FIGS. 67B-67H. Theengine controller 13660 itself, in some embodiments, may be separatefrom but connected to and in communication with a power electronicssoftware and hardware scheme which facilitates conversion of mechanicalto electrical energy essentially downstream from the Stirling. Someembodiments of the power electronics may be those described in U.S.patent application Ser. No. 13/447,897 filed on Apr. 16, 2012 andentitled “MODULAR POWER CONVERSION SYSTEM, which is hereby incorporatedherein by reference in its entirety. While the two systems may sharedata and communications of numerous system variables in someembodiments, the engine controller 13660 is generally understood as aseparate system from the power electronics. The engine controller 13660retains responsibility for all aspects of the Stirling engineoperational control including, but not limited to, regulation of theairlock as described above as well as for example the four burners 13674in the current embodiment of the Stirling engine. Clearly a controllersuch as described here can accommodate control of other Stirling designsas well.

In some embodiments, and as shown in FIGS. 67B-67H, a fuel source 13662provides fuel to each of four separate gas trains including valves 13666and variable flow element 13668 through a main regulator 13664. In someembodiments, each valve 13666 is a dual gas valve, each shown with acombination regulator. In some embodiments the valves 13666 may beMAXITROL CV-300 valves for example, as manufactured by the MaxitrolCompany, Southfield, Mich., USA or White-Rodgers 36H32-423 valves, forexample, as manufactured by White-Rodgers, Saint Louis, Mo., USA. Thevariable flow element 13668 provides a variable flow resistance to varythe flow of fuel to each burner assembly 12907 independently of the airflow through the burner assembly 12907. In some embodiments, themodulating valve may be a MAXITROL EXA-40 as manufactured by theMaxitrol Company, Southfield, Mich., USA. Another embodiment may forgothe main regulator 13664 and regulate the gas pressure with acombination valves 13666 such the MAXITROL CV-300 as manufactured by theMaxitrol Company, Southfield, Mich., USA. In various embodiments, thefuel delivery may be controlled in part by a variable flow element 13668which in turn may be controlled and may be monitored by the enginecontroller 13660 over four respective control signal lines 13670. It isto be appreciated that in some embodiments, the variable flow elementmay be a rotary actuator and a throttle plate. In some embodiments, theactuator 13668 may be a modulating valve such as a MAXITROL EXA-40 asmanufactured by the Maxitrol Company, Southfield, Mich., USA. A blower13672 provides the air flow for combustion in the burner(s) 13674, aswell as cooling of a burner enclosure. The engine controller 13660controls the air flow via a speed command 13678 passed to the variablefrequency drive 13676. A blower speed signal 13677 may provide afeedback signal to the controller 13660 which permits, amongst other,evaluation and control of the blower drive by the engine controller13660. The blower 13672 is shown here as a single blower but may also bea plurality of blowers. A main valve enable relay 13680 and pressureswitch 13682 to enable/disable the fuel valves 13666 is provided inconjunction with the blower 13672 and with a series of heater headtemperature sensors 13685 and a Programmable Logic Controller (PLC)13684. An oxygen sensor 13690 as discussed previously in thisapplication may also be provided in the burner as well to communicateoxygen data to the controller which may also facilitate the fuel andignition control.

In some embodiments, for each of the four burners 13674 a flamedetection sensor 13692 is provided which, as described previously inthis application, is critical for safety, temperature control and theignition process among other things. In some embodiments, the ignitorcircuit 13694 for each of the burners are directly influenced by theflame sensors 13692 through the controller 13660 and the ignitor circuit13694 are controlled via the igniter signal lines 13696 based on theflame sensor data and other data from the engine such as oxygen sensorsfor example. In various embodiments, ignition may be accomplished byeither spark or hot surface ignitors and flame detection sensors mayinclude a flame rod and flame rectification, as well as optical sensingof the flame or alternatively other methods of known flame detection.

In some embodiments, the ignitor circuits 13694 are commercialcombustion control circuits that open the fuel valves and attempt toignite a flame, then monitor the flame, attempt to relight the flame ifit fails and closes if the fuel valves if unable to establish a flamewithin a given amount of amount of time. In some embodiment, thecombustion control circuits are customized variants of Series 35-53 madeby Fenwal Controls of Ashland Mass. In these embodiments, the enginecontroller 13660 enables the ignitor circuit 13694, which each close arelay to power the associated fuel valves 13666, and initiate anignition sequence. In some embodiments, if at any time the ignitorcircuit 13694 is unable to ignite and detect a flame, it will open thefuel valve relay, thereby closing the fuel valve 13666 for that burner.In some embodiments, if the pressure switch 13682 does not detect airflow, the system may interrupt power to all the fuel valves 13664thereby ending combustion and preventing a safety hazard. In someembodiments, if the overtemp circuit 13684 detects excessivetemperatures in a given heater head, then the fuel valve 13666associated with that heater head may be closed to prevent damage to thatheater head and allow it to cool.

In various embodiments, coolant flow and temperature are also inputs tothe controller 13560 to control the coolant flow pump 13698 and ensurethat appropriate coolant temperature is maintained in the Stirlingcycle. In some embodiments, the Airlock delta Pressure Regulator (AdPR)13611 is also directly connected and controlled via the enginecontroller 13660. In some embodiments, the engine controller 13660receives the airlock pressure data from the AdPR 13611 as describedabove and activates the pump in the AdPR to maintain the appropriatepressure differential between the crankcase and the airlock.

In some embodiments, the engine controller 13660 may also communicatewith the power electronics (not shown) over CAN bus but could also, insome embodiments, rely on wireless communications or othercommunications protocols such as USB. The engine controller 13660 may,in some embodiments, command the speed of a permanent magnet synchronousmotor (“PMSM”) motor. Embodiments of power electronics as they relate tocontrol and monitoring of the PMSM motor may be those described in U.S.patent application Ser. No. 13/447,897 filed on Apr. 16, 2012 andentitled “MODULAR POWER CONVERSION SYSTEM, which is hereby incorporatedherein by reference in its entirety. For purposes of this discussionwith regards to one embodiment discussed in the present application, theengine controller 13660 and power electronics may exchange data andcommands including, but not limited to, motor drive velocity command,generator velocity, Bus voltage, Bus current, motor drive IGBT bridgetemp., shunt control, shunt active, battery voltage, batterytemperature, inverter power, inverter enable, inverter PWM, invertervoltage inverter current, inverter temperature, converter power,converter enable, converter PWM, converter voltage, converter currentand converter temperature. Here converter refers to one or more DC/DCconverter circuits. Certain direct inputs into the engine controller13660 may also be necessary and can include but are not limited to oiltemperature from the crankcase, battery temperature, motor temperatureand shunt temperature.

Annular Venturi Burner

FIGS. 70A-70D disclose a further embodiment of an annular-venturi burner13801 for use in conjunction with a multiple heater head and pistonengine as described previously in FIGS. 54-62. This embodiment of theburner 13801 is also specifically directed to the independent heating ofmultiple heater heads 13803. In this embodiment there are four (4)heater heads 13803 and respective burner assemblies 13807, although onlytwo (2) are visible in the cross-section of FIG. 70A, and more or lessheater heads are of course also possible for the engine. As in thepreviously discussed embodiments the heater heads 13803 and burnerassemblies 13807 are encompassed by a burner housing 13811 and eachheater head 13803 is heated by an individual burner assembly 13807. Theburner assembly 13807 is supplied a fuel/air mixture for combustion viaa blower (not shown) that supplies air through the air inlet 13823 and afuel system (not shown) that provides fuel through the fuel inlets13816. The flame may form within the venturi body 13847 and/or in thecombustion chamber 13831 formed by an annular arrangement of the heatertubes 13809. The hot combustion gases then flow past the heater tubes13809 before entering a recuperative preheater 13851-13855. Burnerfairings 13808 in the form of metal rings mounted on a mid-burner plate13805 direct the hot combustion gases across the heater tubes 13809 anddiverts the hot combustion gases from exiting axially from the heatertubes 13809. As described previously in relation to FIGS. 59 and 62 theexhausting combustion gases pre-heat the incoming air in therecuperative heat exchanger, 13851-1385 and then exit the burner housing13811 through exhaust outlets (not shown).

More specifically, observing FIG. 70A, The burner housing 13811 iscomprised of an outer wall 13811A and a manifold 13811B that connectsthe inlet 13823 port to the air side 13852 of the recuperative heatexchanger and the exhaust port to the exhaust side 13852A of therecuperative heat exchanger. In some embodiments, the manifold 13811Bmay include an interface 13804A to the stirling engine. The combustiongases exiting from the heater heads 13803 may be sealed at the interfacewith o-rings or other airtight seals and the manifold 13811B may bemechanically attached to the cooler plate 13810 of the Stirling enginewith a marmin clamp. In other embodiments, the burner 13801 may bemechanically attached to the Stirling engine with a bolted flange orother mechanical means. The heater heads 13803 themselves may be any ofthe various embodiments of tube heater heads described in the precedingsections, including, but not limited to, straight tube heater heads orhelical tube heater heads as disclosed in U.S. patent application Ser.No. 13/447,990, filed Apr. 16, 2012, and incorporated herein. By way ofexample, the present embodiment is contemplated utilizing heater tubes13809 through which flow the working gas, for example helium, which isheated by the respective burner assemblies 13807.

In various embodiments of the burner 13801, the single blower (notshown) may be incorporated to maintain a consistent average air flowsupplied to the burner 13801 and hence to each of the individual burnerassemblies 13807. In some embodiments, the blower may also produce avariable air flow when necessary to control the fuel/air mixture in theventuri 13841. The blower may supply combustion air at a desiredvelocity depending on instructions from a controller for purposes ofignition, then once ignition has occurred, the desired air flow rate maybe regulated by the controller dependent on data received from sensors,which may include, but not limited to, an oxygen sensor sampling thecooled exhaust gas. In some embodiments, a fuel system with variableflow valves for each heater head, which, in some embodiments, may be aMaxitrol EXA-40 valves for example, as manufactured by the MaxitrolCompany, Southfield, Mich., USA, may control the fuel flow to achievethe commanded temperature on each heater head. The blower may beadjusted to achieve the desired fuel/air ratio.

An important aspect of the present embodiment is the efficient heatingof the incoming air through the extraction of waste heat in the exhaustto raise the incoming air temperature thereby improving the efficiencyof combustion processes and the burner 13801. The blower connectsthrough the air inlet 13823 in the outer wall 13811A of the burnerhousing into an air channel 13851. The air channel 13851 extendscircumferentially around the burner inside the outer wall 13812 of theburner 13801 and directs the air developed by the blower B across anintermediate baffle 13853 and up into the hot manifold 13857 beforeentering the swirlers 13882 of the venturi ejectors 13841. Theintermediate baffle 13853 directly separates the incoming air from theexhaust gases exiting the burner through an exhaust channel 13855 andprovides for the heat transfer from the exiting exhaust to the incomingair in the air channel 13851. The heat transfer efficiency across theintermediate baffle 13853 is critical because the hotter the incomingair can be heated, the less fuel is necessary to reach the desiredcombustion temperatures.

The incoming air is preheated to a desired temperature, for example, butnot limited to 600-750° C. These are many reasons preheating theincoming air may be beneficial and these include, but are not limitedto, facilitating ignition and combustion as the air is directed to theburner assemblies 13807 and/or increasing the thermal efficiency of theburner by capturing some of the thermal energy in the combustion gasesexiting the heater heads. In some embodiments, preheating of the air mayreduce the hot exhaust temperature from 900° C. to 300° C. In someembodiments, the amount of preheating which may be accomplished may berelated to the efficiency of heat transfer from the exiting exhaust tothe incoming air. The heat transfer across the intermediate baffle 13853may be improved by adding rows of folded fins 13852 on the air side andfolded fins 13852 a on the exhaust side of the intermediate baffle13853. The folded fins may be brazed to the intermediate baffle 13853 toassure good thermal attachment. In various embodiments, the materialproperties of the folded fins may be optimized for the operatingtemperature. For example, in some embodiments, the rows of folded finsnear the top may be heat resistant metals, which may include, but is notlimited to, INCONEL 625, while lower and cooler folded fins may havehigher thermal conductivity but lower operating temperature. In variousembodiments, the materials for these folded fins may include, but is notlimited to, stainless steel 409 or Ni 201 for example. The preheated airexits the air channel 13851 and is directed radially into a hot airmanifold 13857 which communicates with each of the multiple burnerassemblies 13807 specifically directing the preheated air to the swirlerportion 13882 of the venturi 13841. In various embodiments, thepreheated air enters the hot air chamber 13857 through a substantially360 degree circumferential opening around the exit of the air channel13851. In some embodiments, this may result in a consistent flow rate ofpreheated air delivered to each of the burner assemblies 13807. Invarious embodiments, additional channels or passageways (not shown) maybe provided in the hot air chamber 13857 to direct the preheated air inthe hot air chamber to a specific burner head. In various embodiments,the 360 degree output from the air channel 13851 is used when there isonly one blower B developing the air flow into the engine.

The burner assembly 13807 may be further understood by referring toFIGS. 70B-70D. Several elements of the burner assembly 13807 are shownin FIG. 70B include a venturi body 13841, a fuel inlet 13816, an airswirler 13882, an ignitor 13818, and a flame detector 13860. The venturi13841 in this embodiment is a venturi type ejector as disclosed forexample in U.S. patent application Ser. No. 12/829,320 filed Jul. 1,2010, now U.S. Publication No. US-2011-0011078-A1 published Jan. 20,2011 and entitled Stirling Cycle Machine (Attorney Docket No. 178),which is hereby incorporated herein by reference in its entirety. Insome embodiments the venturi 13841 may be beneficial for many reasons,including, but not limited to, providing the benefit of reducing oreliminating the need for a completely separate fuel control scheme asregulation of the airflow changes the vacuum which, in turn,correspondingly affects the fuel flow and regulates the burner power. Insome embodiments, the venturi allows use of typical gas pressures inbuildings, e.g., 7 inches of water column, without a compressor. Theblower forces air through radial swirler vanes 13882. The flow ofswirling air mixes with fuel in the venturi throat 13847 and froms aswirl stailized flame in the expanding section of the venturi 13849and/or in the combustion chamber 13831 of FIG. 70C.

Referring again to FIG. 70B, the ignitor 13818 ignites the fuel-airmixture in the venturi throat 13847. The ignitor 13818 may be a sparkplug and may function as the high voltage electrode of a flameionization flame detector. In other embodiments, the ignitor 12918 maybe a hot surface ignitor. In one embodiment the ignitor are SiliconNitride Hot Surface ignitors produced by Crystal-Technica of SouthGrafton, Mass. The ignitor 13818 may located advantageously near thecenter of the venturi 13841 to promote a uniform composition and flowfrom the venturi 13841. The ignitor 13818 may be mounted in the ignitorport 13817B.

The flame detector 13860 mounts in the 13817C port and provides a signalto the controller indicative of the presence or absence of a flame fromthe burner assembly 13807. In one embodiment, the flame detector 13860comprises a temperature sensor 13862 inside a heat resistant tube 13861.The temperature sensor 13862 may be a type K thermocouple in an inconelsheath, a type B or R thermocouple or other high temperature sensor. Theheat resistant tube 13861 may be a heat resistant metal such as inconel625, Mar-M or it may be a ceramic tube formed from zirconia or otherhigh temperature ceramics. In other embodiments the port 13817B may beused for a vision based flame detection circuit including but notlimited to one or more of the following: IR flame detector, visiblelight flame detector and/or UV flame detector. In another embodiment,the flame sensor may be a flame rod connected to a Series 35-52 Ignitioncontroller as manufactured by by Kidde-Fenwal Inc. of Ashland Mass.

Still referring to FIG. 70B, the fuel enters the burner assembly 13816through the fuel inlet 13816, that is mounted in port 13817A. The fuelflows into a plenum or manifold 13871 before flowing through the fuelports into the venturi throat 13847. The mixing of the fuel and air inthe venturi throat is best shown in FIG. 70C, which is a detailed viewof the venturi throat and fuel injectors. The air enters the venturithroat 13847 with an induced swirl created by radial vanes 13882 thatare described in detail below. The air initial flows in radially towardthe ignitor 13818 and then is directed into an axial flow by anaxisymetric protrusion 13880 around the ignitor and the inlet to theventuri 13877A. In one embodiment pictured in FIG. 70C, the inlet to theventuri 13877A is part of the venturi bushing 13877. In otherembodiments the inlet 13877 may be an integral part of the venturi body13841 or the a conical or a separate piece. In some embodiments, theinlet 13877A and the protrusion 13880 may be formed to provide anapproximately constant cross-section flow area as the flow changes fromradial to axial. One possible benefit of an approximately constant flowarea is a minimization of pressure drop through the burner assembly13807. The protrusion 13880 may be conical or may have a surface with anincreasing slope from near horizontal to approaching vertical.

Still referring to FIG. 70C, the fuel enters the fuel plenum 13873 fromthe fuel inlet tube 13816 that is connected to the fuel system that wasdescribed above in reference to FIG. 62. The fuel plenum forms anannular space around the outside diameter of the venturi body 13841 andsupplies fuel to a plurality of fuel ports 13875. The fuel ports 13875provide fuel to an annular space between the venturi bushing 13877 andthe venturi body 13841. This annular space is mostly filled with fueland is herein referred to as the fuel annulus. The fuel exits the fuelannulus flowing axially along the walls of the venturi throat 13847. Theair flowing axially through center of the bushing 13877 and into theventuri throat 13847 creates a region of low pressure downstream of thefuel annulus that draws the fuel into venturi throat 13847 where itmixes with air. The exit of the fuel annulus is preferably upstream ofthe ignitor 13818 by a sufficient distance to create an ignitablemixture of fuel-air next to the hot surface or spark of the ignitor13818. The fuel annulus may have a very thin annular opening to maximizefuel flow velocities. In other embodiments, the annulus is larger. Ingeneral, the annular exit of the fuel annulus has a constant radial gapto maximize fuel flow uniformily around the venturi throat 13847. In apreferred design, the gap is 1/20 of the venturi throat diameter or hasa radial gap of 0.035″. The fuel ports 13875 may be radial or may enterthe fuel annulus at an angle to induce a swirl in the fuel.

FIG. 70D presents an isometric view of the swirl vane plate 13882mounted on the venturi body 13841, where the venturi entrance 13877A isvisible. The radial vanes 13882A impart a tangential velocity or swirlmotion to the radially flow air. In one embodiment, the radial vanes13882A are straight and not curved. In another embodiment the vanes arecurved so that the vanes are radial at the outside diameter of theradial vane plate, the vanes curve until the flow leaves the inward edgeof the vane with the desired swirl. In FIG. 70D, the vanes are curvedand aerodynamic to minimize pressure drop through the burner. The vanesare essentially airfoils that are initially thick near the outsidediameter of the radial venturi plate and increasingly thin toward theventuri inlet.

One theory on the advantage of injecting the fuel through and the fuelannulus is that the annulus provides a more uniform provision of fuelaround the venturi throat by providing a plenum for the fuel from theplurality of fuel jets to mix and flow uniformly into the venturi throat13847. Another theory on the advantage of the fuel annulus is that itadvantageously places the fuel next to the wall, where the local airflow may be more uniform than in the center. Still another theory isthat the lack of fuel jets across the airflow avoid disturbing theairflow and result in a more uniform air flow exiting from the venturithroat 13847.

While the principles of the invention have been described herein, it isto be understood by those skilled in the art that this description ismade only by way of example and not as a limitation as to the scope ofthe invention. Other embodiments are contemplated within the scope ofthe present invention in addition to the exemplary embodiments shown anddescribed herein. Modifications and substitutions by one of ordinaryskill in the art are considered to be within the scope of the presentinvention.

What is claimed is:
 1. An external combustion engine containing aworking fluid and comprising: a burner element for heating the workingfluid of the engine; at least one heater head defining a working spacecontaining the working fluid; at least one piston cylinder containing apiston for compressing the working fluid; at least one cooler forcooling the working fluid; a crankcase comprising: a crankshaft forproducing an engine output; a piston rod connected to the piston; adrive mechanism that converts the linear motion of the piston rod torotary motion of the crankshaft; and a linear cross-head bearingcomprising a journal and a guide, wherein the journal has a first endrigidly attached to the piston rod and a second end attached to thedrive mechanism, wherein the guide is located outside the working space,and wherein the linear cross-head bearing solely constrains the motionof the piston.
 2. The external combustion engine of claim 1, wherein theratio of the linear cross-head bearing length over diameter is greaterthan 2.0.
 3. The external combustion engine of claim 1, wherein thelinear cross-head bearing is a hydrodynamic bearing supplied withlubricating fluid from an annulus on the guide.
 4. The externalcombustion engine of claim 1, further comprising a rod seal assemblythat comprises: a housing between two spaces configured to receive thepiston rod, the piston rod disposed between the crankcase and theworking space; a floating clearance bushing configured to move axiallyand radially within the housing and disposed coaxially around the pistonrod and forms a clearance seal with the piston rod; and at least onestationary annular element fixed within the housing configured to form aface seal with the floating clearance bushing.
 5. The externalcombustion engine of claim 1 further comprising a rod seal assembly thatcomprises: a housing between two spaces configured to receive the pistonrod, the piston rod disposed between the crankcase and the workingspace; floating rod seal assembly comprising at least one rod sealmounted onto the floating rod seal assembly.
 6. The external combustionengine of claim 5, wherein the rod seal is a spring energized seal. 7.The external combustion engine of claim 5, the floating rod sealassembly further comprising: an outer ring; and at least one bushing. 8.The external combustion engine of claim 5, the rod seal assembly furthercomprising: a scraper ring; a particle trap; a port; and a filter. 9.The external combustion engine of claim 1 wherein the piston comprises aclearance seal.
 10. The external combustion engine of claim 9, whereinthe clearance seal is a cylinder with an inner diameter facing thepiston and an outer diameter facing the cylinder, wherein the outerdiameter includes a plurality of grooves on the outside diameter and thegrooves are perpendicular to the piston axis.
 11. The externalcombustion engine of claim 9, wherein the clearance seal is a cylinderwith an inner diameter facing the piston, wherein at least one radialseals touches the inner diameter face and the piston.
 12. The externalcombustion engine of claim 9, wherein the clearance seal comprises atleast one of the following materials: graphite, PTFE, UHMWPE, antimony.13. The external combustion engine of claim 9, wherein the clearanceseal comprises at least graphite and antimony.
 14. An externalcombustion engine containing a working fluid and comprising: a burnerelement for heating the working fluid of the engine; at least one heaterhead defining a working space containing the working fluid; at least onepiston cylinder containing a piston for compressing the working fluid;at least one cooler for cooling the working fluid; a crankcasecomprising: a crankshaft for producing an engine output; a piston rodconnected to the piston; a rocking beam driven by the piston rod; aconnecting rod connected at a first end to the rocking beam and at asecond end to a crankshaft to convert rotary motion of the rocking beamto rotary motion of the crankshaft; and a linear cross-head bearingcomprising a journal and a guide, wherein the journal has a first endrigidly attached to the piston rod and a second end attached to thedrive mechanism, wherein the guide is located outside the working space,and wherein the motion of the piston is solely constrained by the linearcross-head.
 15. The external combustion engine of claim 14, wherein theratio of the linear cross-head bearing length over diameter is greaterthan 2.0.
 16. The external combustion engine of claim 14, wherein thelinear cross-head bearing is a hydrodynamic bearing supplied withlubricating fluid from an annulus on the guide.
 17. The externalcombustion engine of claim 14, further comprising a rod seal assemblythat comprises: a housing between two spaces configured to receive thepiston rod, the piston rod disposed between the crankcase and theworking space; a floating clearance bushing configured to move axiallyand radially within the housing and disposed coaxially around the pistonrod and forms a clearance seal with the piston rod; and at least onestationary annular element fixed within the housing configured to form aface seal with the floating clearance bushing.
 18. The externalcombustion engine of claim 14, further comprising a rod seal assemblythat comprises: a housing between two spaces configured to receive thepiston rod, the piston rod disposed between the crankcase and theworking space; floating rod seal assembly comprising at least one rodseal mounted onto the floating rod seal assembly.
 19. The externalcombustion engine of claim 18, wherein the rod seal is a springenergized seal.
 20. The external combustion engine of claim 18, thefloating rod seal assembly further comprising: an outer ring; and atleast one bushing.
 21. The external combustion engine of claim 18, therod seal assembly further comprising: a scraper ring; a particle trap; aport; and a filter.
 22. The external combustion engine of claim 14,wherein the piston comprises a clearance seal.
 23. The externalcombustion engine of claim 22, wherein the clearance seal is a cylinderwith an inner diameter facing the piston and an outer diameter facingthe cylinder, wherein the outer diameter includes a plurality of grooveson the outside diameter and the grooves are perpendicular to the pistonaxis.
 24. The external combustion engine of claim 22, wherein theclearance seal is a cylinder with an inner diameter facing the piston,wherein at least one radial seals touches the inner diameter face andthe piston.
 25. The external combustion engine of claim 22, wherein theclearance seal comprises at least one of the following materials:graphite, PTFE, UHMWPE, antimony.
 26. The external combustion engine ofclaim 22, wherein the clearance seal comprises at least graphite andantimony.
 27. An external combustion engine containing a working fluidand comprising: a burner element for heating the working fluid of theengine; at least one heater head defining a working space containing theworking fluid; at least one piston cylinder containing a piston forcompressing the working fluid; at least one cooler for cooling theworking fluid; a crankcase comprising: a crankshaft for producing anengine output; a piston rod connected to the piston; a drive mechanismthat converts the linear motion of the piston rod to rotary motion ofthe crankshaft; and a piston comprising: a piston base wherein at leastone piston seal is mounted; an outer shell that mounts to the base anddefines the outside of the piston above the base; and an inner shell hasis shorter and narrower than the outer shell and defines with the pistonbase a volume above the inner surface of the piston base.
 28. Theexternal combustion engine of claim 27, wherein the piston basecomprises a small orifice to allow the pressure inside the piston toequalize with average pressure of the work space.
 29. The externalcombustion engine of claim 28, wherein the orifice has diameter between0.002 and 0.008 inches.
 30. The external combustion engine of claim 28,wherein the inner shell comprises a port to allow fluid communicationbetween the volume inside the inner shell and the volume between theinner and outer shells.